Control system for controlling engagement of an automatic transmission torque converter clutch

ABSTRACT

An electronically controlled multiply ratio transmission for an automotive vehicle powertrain having a hydrokinetic torque converter with a bypass clutch, solenoid operated shift valves under the control of an electronic microprocessor that receives signals from an output shaft speed sensor, an engine speed sensor, a driver controlled drive range selector position, an engine throttle position sensor, and other powertrain operating variables, the bypass clutch having a controlled slip characteristic.

Cross-Reference To Related Application

This application is a division of U.S. Ser. No. 07/926,627, filed Aug.10, 1992, entitled "AUTOMATIC TRANSMISSION CONTROL SYSTEM", now U.S.Pat. No. 5,305,663.

TECHNICAL FIELD

Our invention relates to automotive vehicle drivelines and particularlyto improvements in a multiple ratio hydrokinetic transmission mechanismwith a hydraulic control valve system that is under the control of anelectronic microprocessor.

BACKGROUND OF THE INVENTION

Our invention relates to automatic controls for effecting ratio changesin a hydrokinetic power transmission mechanism of the kind disclosed inU.S. Pat. No. 4,934,216, which is assigned to the assignee of thepresent invention. The transmission of the '216 patent includes ahydrokinetic torque converter having an impeller and a turbine situatedbetween an engine crankshaft and a compound planetary gear system, thelatter having torque input elements adapted to be clutched to theturbine of the torque converter and output elements adapted to beconnected to the traction wheels through a differential and axlemechanism. The transmission of the '216 patent is adapted particularlyto be used in a vehicle driveline having a forward engine and rearwardtraction wheels. It is contemplated, however, that the principles of thepresent invention can be applied to a transaxle transmission of the typedisclosed in U.S. Pat. No. 5,081,886, which also is assigned to theassignee of this invention.

Other examples of prior art transmissions capable of being adapted toincorporate the improvements of our invention are described in U.S. Pat.Nos. 4,978,328 and 5,083,481. These also are assigned to the assignee ofthis invention.

The transmission of the present invention comprises fluid pressureoperated clutch and brake servos that control the relative motion of theelements of the planetary gearing to establish four forward drivingratios, including an overdrive ratio and a single reverse ratio. Ahydrokinetic torque converter, which forms a hydrokinetic torque flowpath from the engine crankshaft to the input elements of the gearing,includes a turbine and an impeller arranged in a toroidal fluid flowcircuit. It includes also a friction bypass clutch adapted to connectthe impeller to the turbine to establish a mechanical torque flow pathin parallel with respect to the hydrokinetic torque flow path of theconverter.

Prior art U.S. Pat. No. 5,029,087 discloses an electronic controlstrategy for effecting a controlled slip in a torque converter bypassclutch whereby the bypass clutch is actuated by modulated converterclutch solenoid pressure from a clutch solenoid valve to effect varyingclutch capacity so that the resulting controlled slip results in anactual slip that approaches a target slip determined by the operatingparameters of the driveline. That controlled slip strategy has featuresthat are common to the control strategy for the bypass clutch of thepresent invention, as will be explained.

In that prior art bypass clutch design, the actual converter slip iscontinuously monitored by the processor as the engine speed and theturbine speed are detected by engine speed and turbine speed sensors. Adesired slip is calculated by the processor during each backgroundcontrol loop and that value is subtracted from the actual slip to detecta slip error. The slip error, in turn, is used to calculate a duty cyclefor the clutch solenoid valve so that the error is reduced and thedesired slip, together with the actual slip, approach a target value.The magnitude of the target value is a value stored in computer memory.The target slip that is fetched from memory depends upon the operatingvariables of the driveline. In this manner, the desired slip and theactual slip approach the target value asymptotically.

The bypass clutch strategy of such prior art designs will provide for acontinuous slip to eliminate transient torque fluctuations and noise,vibration and harshness during steady-state operation. Thenon-conformance of the desired slip to the target value due to theasymptotic approach of the desired slip with respect to the target valuefetched from memory tends to make calibration of the driveline,particularly the bypass clutch, more difficult. The precise desired slipthat is required by the processor to satisfy a particular drivelinecondition is not available because the desired slip versus timerelationship will float: relative to the target value during itsasymptotic approach to the target. Furthermore, the continuous slippingthat is characteristic of such bypass clutch strategy during steadystate operation may cause bypass clutch friction surface durabilityconcerns.

BRIEF DESCRIPTION OF THE INVENTION

The improved bypass clutch control strategy of the present inventionovercomes possible shortcomings, described above, with respect to theprior art control strategy for a modulated bypass torque converterclutch wherein the actual slip or the desired slip approach a targetslip value asymptotically. The improved control strategy of the presentinvention makes it possible to compute the desired slip in a bypassclutch controller of the kind previously described so that at thebeginning of the bypass clutch operating mode, the desired slip is madeequal to the actual slip that is measured by the engine speed sensor andthe output shaft speed sensor or turbine speed sensor. Having determinedthe desired slip by setting it equal to the actual measured slip, a rampmodifier is fetched from the computer memory so that the desired slipwill decrease over time at a prescribed rate that depends upon the valuefor the modifier that is fetched from memory. That modifier is dependentupon the gear ratio that is called for by the transmission shift controlstrategy. Thus, the slope of the ramp for the desired slip over timewill be a unique ramp for each gear ratio. The desired slip then willapproach the target value until it matches the target value, the latterbeing fetched from memory in the same manner described above withrespect to the control system of prior art U.S. Pat. No. 5,029,087.

After the desired slip and the target value are equal, one with respectto the other, the actual slip then will continue to decay until it tooreaches the target value. It continues to decay toward a valuecorresponding to the target value as long as there is an error in theslip as measured by the difference between the actual slip and thedesired slip. When the speed and load conditions dictate, the targetslip value becomes zero and the control strategy will initiate ahardlock mode whereby the slip will be ramped again down to a zero slipusing the same control technique that is used to determine desired slipat the initiation of the engagement of the bypass clutch. That is, aslip modifier value again is fetched from the ROM portion of the memory,the value of the modifier depending upon the gear ratio in place. Whenthe converter is in the hardlock mode, a direct mechanical connection isestablished between the engine and the torque input element or turbineshaft.

Provision is made also in the improved control strategy for the bypassclutch of the present invention for interrupting the hardlock modeduring ratio shifts. The beginning of the shift is detected by measuringa change in speed ratio. This signals the interruption of the hardlockmode and the beginning of an open loop converter control. During thetime interval for the open loop converter control, the converter clutchactuating pressure is reduced by a multiplier fetched from memory,thereby reducing the torque transmitting capacity of the clutch so thattransient torque fluctuations in the driveline are absorbed by theconverter during a shift interval. The termination of the shift isdetected by continuously monitoring the speed ratio. When the speedratio change (delta speed ratio) is large enough to indicate the end ofthe shift, the control strategy will cause the bypass clutch for theconverter to reenter the hardlock mode or the slipping, closed loopcontrol mode, as appropriate. This occurs as the absolute slip becomesstabilized at zero RPM. After that occurs, the duty cycle again isramped up to its maximum value corresponding to full bypass clutchengagement if the controller calls for operation of the clutch in thehardlock mode.

The bypass clutch for the hydrokinetic torque converter is capable ofbeing modulated to effect a controlled slip, as explained above, beforethe "hardlock" mode occurs. This is done, as explained, by means of abypass clutch control valve under the control of a modulated converterclutch solenoid valve.

The duty cycle of the modulated converter clutch solenoid valve isdetermined by a microprocessor so that the absolute slip for anyoperating condition will approach over time the target slip that isobtained from the memory portion of the microprocessor to satisfy theinstantaneous operating demands. This tends to develop heat because ofthe energy loss, during the slipping mode. If the temperature of thetransmission fluid is above a predetermined value, the microprocessorsoftware will trigger a maximum duty cycle for the modulated converterclutch solenoid so that the converter clutch will assume a fully engagedcondition to avoid slipping until the temperature reaches a value lessthan the predetermined maximum.

According to another feature of the invention, we have provided asimplified assembly procedure for assembling the valve body to thetransmission casing which includes the use of dowel bolts that align thegaskets, the cover plates for the valve body and the valve body itselfwith respect to the transmission case to which the valve body issecured, thereby providing precise registry of the stacked elements ofthe control valve body and its associated gaskets and cover plates. Thisimproves the reliability of the assembly and simplifies the assemblyprocedure.

These and other improvements will become apparent from the followingparticular description, which will refer to the figures of the drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a partial cross-sectional view of an automatic transmissionhaving a hydrokinetic torque converter and gear elements capable ofbeing controlled by the improved control valve system of this invention.

FIG. 2 is a full schematic representation of the converter and the gearelements of the cross-sectional view of FIG. 1.

FIG. 3 is a chart that shows the clutch and brake engagement and releasepattern for effecting ratio changes in the transmission of FIGS. 1 and2.

FIG. 3A is a schematic overview of a microprocessor adapted to controlthe valve circuit of the present invention.

FIG. 4 shows in schematic form the elements of a valve circuit for thepresent invention.

FIGS. 4A and 4B show a schematic valve diagram for the control valvesystem shown in FIG. 4 when the transmission is in neutral or park withthe throttle closed and the converter open.

FIGS. 5A and 5B show a schematic valve diagram of the valve system ofthe invention wherein the valve elements are positioned for reversedrive, part throttle, open converter operation.

FIGS. 6A and 6B show a schematic valve diagram of the valve system ofthe invention wherein the valve elements are positioned for neutral,closed throttle, open converter operation.

FIGS. 7A and 7B show a schematic valve diagram of the valve elements inthe positions they assume when the transmission is conditioned foroverdrive, first gear, closed throttle, open converter operation.

FIGS. 8A and 8B show a schematic valve diagram of the valve elements intheir positions corresponding to overdrive, second gear, part throttle,open converter operation.

FIGS. 9A and 9B show a valve circuit diagram with the valve elements inthe positions they assume during overdrive, third gear operation withthe bypass clutch control valve in the position it assumes when theconverter is in its slipping mode. The 2-3 back-out valve positionmainly depends upon whether the throttle is open or closed.

FIGS. 10A and 10B show a schematic valve circuit diagram with the valveelements in the positions they assume during overdrive, fourth gear,part throttle operation with the torque converter clutch applied. Thebypass clutch control valve is shown in its position when the converterclutch is applied and the converter is in a lock up mode.

FIGS. 11A and lib show a schematic valve circuit diagram with the valveelements in the positions they assume when the transmission is in theD-range, third gear operation with part throttle and with the converterin the modulating mode.

FIGS. 12A and 12B show a schematic valve diagram with the valve elementsin the positions they assume when the manual valve is in the manual lowposition and the transmission is conditioned for first gear operationwith closed throttle and an open converter.

FIGS. 13A and 13B show a schematic valve diagram with the valve elementsin the positions they assume during operation in the manual low driverange when in the second gear, part throttle, open converter operatingmode.

FIGS. 14A and 14B show a schematic valve diagram with the valve elementsin the positions they assume during forward drive engagement of theforward clutch with a closed throttle.

FIGS. 15A and 15B show a schematic valve diagram with the valve elementsin the positions they assume during reverse drive as the low and reverseclutch becomes applied with a closed throttle.

FIGS. 16A and 16B show a circuit diagram with the valve elements in thepositions they assume during overdrive operation and during a 1-2upshift with part throttle.

FIGS. 17A and 17B show a schematic valve diagram with the valve elementsin the positions they assume during overdrive operating and during a 2-3shift with part throttle.

FIGS. 18A and 18B show a valve diagram with the valve elements in thepositions they assume during overdrive operation and during 3-4 upshiftwith part throttle. The 2-3 backout valve for the 4-3 downshift at partthrottle corresponds to that shown in FIG. 18B.

FIGS. 19A and 19B show a schematic valve diagram with the valve elementsin the positions they assume during overdrive operation with the vehiclecoasting during a 4-3 downshift and with a closed throttle.

FIGS. 20A and 20B show a schematic valve diagram with the valve elementsin the positions they assume during a 4-3 part throttle downshift in theoverdrive operating mode.

FIGS. 21A and 21B show a schematic valve diagram with the valve elementsin the positions they assume during a 2-1 closed throttle downshift inthe overdrive operating mode.

FIGS. 22A and 22B show a schematic valve diagram with the valve elementsin the positions they assume during a 3-2 closed throttle downshiftduring operation in the overdrive mode.

FIG. 22C is an exploded assembly view showing the valve body, the oilpan, the gaskets and the dowel screws used during the assembly of thevalve body elements to the transmission case.

FIG. 22D is a plan view showing the valve cover plate assembled to thevalve body and located by the dowel pins illustrated schematically inFIG. 22B.

FIG. 22E is a detailed view of the dowel pin or bolt used in theassembly of the valve plates, the gaskets and the valve body of FIG.22C.

FIG. 22F is an end view of the valve body with the valve cover platesapplied and the dowel bolts in place.

FIG. 23 is a chart showing the solenoid states for effecting ratiochanges and the condition of the friction clutches and brakes duringeach of the drive ranges and for each of the ratios in each drive range.

FIG. 24 is a chart showing the automatic transmission shift schedule,which is a plot of the throttle opening versus vehicle speed.

FIG. 24A is a chart showing the relationship between the throttlepressure made available to the 2-3 back-out valve throughout a range ofthrottle positions for both power-on and power-off modes.

FIG. 24B is a flow chart showing the electronic strategy for triggeringthe operation of the 2-3 back-out valve and the orifice control valveduring forward clutch engagements.

FIG. 24C is a chart showing the flow patterns through the variousorifices that are established by the operation of the 2-3 back-out valveand the orifice control valve.

FIG. 25 is a cross-sectional view showing the electronic pressurecontrol solenoid valve and the solenoid actuator which controls the linepressure maintained by the pressure regulator.

FIG. 26 is a cross-sectional view showing the modulated converter clutchcontrol solenoid valve for controlling the torque converter clutch.

FIG. 26A is a flow diagram of the electronic control strategy for thetorque converter bypass clutch during PID closed loop clutch control.

FIG. 26B is a chart showing the relationship between absolute slip,desired slip and target slip for the bypass clutch.

FIG. 26C is a chart that illustrates the relationship between duty cyclefor the electronic modulated converter clutch control solenoid valve,engine speed and speed ratio during a ratio shift.

FIG. 26D shows the relationship between the speed ratio and duty cycleduring the period of time following the command of a ratio shift to theend of that shift.

FIG. 26E is a flow chart showing the electronic control strategy for thehardlock mode for the torque converter bypass clutch following theclosed loop control for the bypass clutch after the absolute slip hasbecome equal to target slip.

FIG. 27 is a cross-sectional view showing the shift control solenoidwhen it is deenergized.

FIG. 27A is a view of the electronic shift control solenoid of FIG. 27with the armature element of the solenoid in the position it assumeswhen the solenoid is energized.

FIG. 28 is a flow chart showing a portion of the electronic controlstrategy for a prior art bypass clutch design.

FIG. 29 is a plot of the bypass clutch slip characteristics for theprior art bypass clutch design to which the chart of FIG. 28 refers.

FIGS. 30 and 30A show a flow chart for the electronic control shiftmodulation logic for the bypass clutch of the present invention when thecontrol system for the transmission is undergoing a ratio change.

PARTICULAR DESCRIPTION OF THE INVENTION The Converter And Gearing

In the cross-sectional view of the transmission seen in FIGS. 1 and 2,numeral 10 designates a hydrokinetic torque converter and numeral 12designates a compound planetary gear unit. The converter 10 and the gearunit 12 are located in a transmission housing illustrated at 14.

The converter 10 includes a bladed impeller 16, a bladed turbine 18, anda bladed stator 20. The converter elements 16, 18 and 20 form a toroidalfluid flow path in known fashion, whereby impeller torque is multipliedhydrokinetically to produce a turbine torque that is distributed throughturbine hub 22 to the turbine shaft 24. The impeller is enclosed withinan impeller housing 26, which is bolted by means of bolts 28 to anengine crankshaft. The bolts 28 are located at the hub of a drive plate30. The latter being secured to the outer periphery of the impellerhousing 26.

Stator 20 is mounted on a one-way brake 32 and is supported bystationary turbine sleeve shaft 34.

Housing wall 36 forms an enclosure for the transmission pump 38, whichincludes positive displacement Gerotor pump elements connected drivablyto the hub 40 of the impeller 16.

A torque bypass clutch generally shown at 42 includes a clutch plate 44adapted to engage the adjacent wall of the impeller housing 26. It issecured to turbine hub 22 by means of a damper assembly 46. Fluid isdistributed radially outward through the space between the clutch plate44 and the adjacent wall of the impeller housing when the clutch isdisengaged. The converter at that time acts as an open converter and iscapable of multiplying torque hydrokinetically. Fluid is suppliedcontinuously to the toroidal cavity of the converter and the pressurethus developed applies the clutch by engaging the clutch plate 44against the adjacent frictional surface of the impeller housing. Theradial outward flow through the space between the plate 44 and theadjacent wall of the impeller housing is interrupted when the clutch isapplied.

The torque delivered to the turbine shaft 24 is transferred through theturbine shaft to the torque input side 48 of reverse clutch 50 and tothe torque input side 52 of forward clutch 54. The output side 56 of theforward clutch 54 is connected to sun gear 58 of the planetary gear unit12. The ring gear 60 of the gear unit 12 is connected to a torque outputshaft 62 through torque transfer member 64.

Sun gear 58 engages a first set of planet pinions 66 supported oncarrier 68. Pinions 66 engage companion pinions 70 which in turn meshwith the ring gear 60. Pinions 70 mesh with a second sun gear 72. Thetorque output side of the reverse clutch 50 is connected to sun gear 72through torque transfer member 74. A brake drum 76 forms a part of thetorque output portion of reverse clutch 50. The brake band for drum 76is applied during overdrive operation to anchor the sun gear 72.

Carrier 68 journals both sets of pinions 70 and 66 and is connected toreverse brake drum 78. Brake band 80 surrounds brake drum 78 and isapplied during reverse drive operation. An overrunning brake 82 anchorsthe carrier 68 during operation in the first speed ratio as forwarddrive reaction torque is delivered to the housing 14.

Carrier 68 is adapted to be connected to the turbine shaft 24 throughdirect-drive clutch 84.

Brake drum 76 is connected to the outer race 86 of an overrunning brake88. Race 86 is adapted to be braked by friction brake 90 to thetransmission housing 14. When brake 90 is applied, overrunning brake 88is adapted to deliver reaction torque to the housing through thefriction brake 90 during intermediate ratio operation.

For a particular description of the mode of operation of thetransmission of FIGS. 1 and 2, reference may be made to U.S. Pat. No.4,934,216 discussed above. For present purposes however, the mode ofoperation can be summarized by referring to FIG. 3. This figure is achart showing the clutches and brakes that are engaged or released toestablished each of the ratios and each of the multiple drive ranges.

To simplify the description, FIG. 3 carries the symbol B1 to identifythe overdrive brake band 76, the symbol B2 to identify the reverse brakeband 80, the symbol C1 to identify the forward clutch 54, the symbol C2to identify the reverse clutch 50, the symbol C3 to identify the directand overdrive clutch 84, the symbol C4 to identify the overrunning brake82, the symbol C5 to identify the intermediate clutch 90 and the symbolC6 to identify the overrunning brake 88. FIG. 3 also shows solenoidstates for solenoid SS1 and solenoid SS2, which will be describedsubsequently.

During operation in the first gear ratio in the automatic drive mode,clutch C1 is applied and brake C4 is applied. Torque delivered to theturbine shaft 24, and then it is transferred through the clutch C1 tothe sun gear 58. The carrier 68 acts as a reaction member since it isbraked by brake C4. Thus, the ring gear 60 is driven in a forwarddriving direction with the highest torque multiplication ratio. Ifcoast-braking is desired (manual range), brake band 80 is applied thuspermitting the reaction torque to be distributed to the housing 14 in areverse driving direction.

An upshift to the second ratio is achieved by applying intermediatebrake C5. This permits the sun gear 72 to act as a reaction point andthe overrunning brake C4 begins to overrun. Torque is distributed to thehousing through the brake C5 and through the overrunning brake C6.

An upshift of the third ratio from the second ratio is achieved byengaging direct-drive clutch C3 while clutch C1 remains applied. Thus,all of the elements of the gearing then are locked together for rotationin unison.

Fourth ratio is achieved by releasing clutch C1 and applying brake bandB1. Sun gear 72 then acts as a reaction point as the input torque isdelivered through the clutch C3, thus overdriving the ring gear 60.

Reverse drive is obtained by applying brake band 80 thus anchoring thecarrier. Engagement of the reverse clutch 50 results in torque transferfrom shaft 24 to the sun gear 72. With the carrier 68 acting as areaction member, ring gear 60 is driven in a reverse direction as sungear 72 acts as a torque input element.

The Microprocessor Controller

FIG. 3A shows a microprocessor that is used to control the valvecircuits that in turn control distribution and exhaust of actuatingpressure to the clutch and brake servos for the transmission. Theprocessor is shown at 92 in FIG. 3A.

As schematically represented in FIG. 3A, an air-charge temperaturesensor 94 is adapted to develop an ambient air temperature that is usedby the processor in developing commands issued to the control valvesystem. The processor also responds to an air conditioning clutch signalfrom sensor 96 which indicates whether the air conditioning system is onor off. This is one of the parasitic torque losses that must beaccounted for by the processor in issuing instructions to the solenoidvalves of the valve circuit.

A brake on/off switch 98 is triggered by the vehicle brakes and theon/off signal is delivered to the processor.

An engine speed sensor 100 measures crankshaft speed and engine coolanttemperature is sensed by temperature sensor 102.

The drive range selected by the operator is indicated by a manual leverposition sensor 104. A transmission output shaft speed sensor 106provides an indication of the driven shaft speed for output shaft 62.That speed is related to the vehicle speed signal developed by sensor108, the vehicle speed being a function of output shaft speed times thegear ratio that exists at any instant. A transmission oil temperaturesignal is delivered to the processor by sensor 110. An engine throttleposition signal is delivered to the processor by sensor 112.

The transmission does not include a turbine speed sensor. However, aturbine speed value readily can be determined by the processor if eitherthe output shaft speed or the vehicle speed is known and if the gearratio is known. These variables that determine speed can be used by theprocessor in computing turbine speed at any instant. The turbine speedthen can be compared to engine speed to detect torque converter slip atany instant.

The signals that are delivered to the valve circuit are received by anelectronic pressure control solenoid 114, which is indicated generallyin FIG. 25. The duty cycle for that solenoid can be changed to developthe required circuit pressure at any instant so that the capacities forthe clutches and the brake servos are maintained at every instant attheir optimum values for the required torque delivery capacity. The dutycycle is determined by the throttle position sensor signal developed bysensor 112 as well as the signal developed by mass air flow sensor 116,which measures the mass air flow at any instant at the engine throttlebody for the internal combustion engine with which the transmission isused. The slip value that is determined by the processor is used todevelop a duty cycle that is received by modulated converter clutchcontrol solenoid 118. This solenoid is shown in FIG. 26.

The control valve circuit includes two solenoid operated shift valveswhich receive shift signals. These are on/off signals from theprocessor. They are received by shift solenoid number 1, shown at 120,and shift solenoid number 2, shown at 122. These shift solenoids will bedescribed subsequently. Solenoid 120 is illustrated in FIGS. 27 and 27A.Solenoid 122 has not been separately illustrated in the drawings sinceit is similar to solenoid 120. Both solenoids are packaged in a singleassembly.

The sensor inputs, such as the engine-related sensor signals indicativeof engine coolant temperature, barometric absolute pressure, etc., areused by the processor to develop more accurate outputs as the load andclimate conditions change. Other inputs are based on driver commandssuch as the engine throttle position. Still other inputs to theprocessor are developed by the transmission itself, such as the outputshaft speed sensor signal, the manual lever position signal and thetransmission oil temperature signal. The processor will develop theappropriate shift time and conditions for shifts in the ratio as well ascontrol the clutch application and release. Line pressure also isdeveloped by the processor to establish optimum shift feel.

The processor is an integrated central processor which converts signals,such as the signals from a vehicle speed sensor and an engine throttleposition sensor, engine temperature sensor, turbine speed sensor and themanual selector lever, into electrical signals for the shift solenoidvalves, the solenoid valve for the converter bypass clutch and thevariable force solenoid for the electronic pressure control. Theprocessor receives the sensor signals and operates on them in accordancewith programmed control strategy, which will be described. The processorincludes appropriate gates and driver circuits for delivering the outputof the operation of the strategy to the hydraulic control valve body forthe transmission.

The processor includes a central processor unit (CPU) having a read onlymemory (ROM) in which the programs (strategy) and calibration data arestored, a control unit that includes a read-write memory or RAM andinternal busses between the memory and the central processor arithmeticlogic unit.

The processor executes programs that are obtained from the memory andprovides the appropriate control signals to a valve circuit as the inputsignal conditioning portions of the processor reads the input data andthe computation logic portions deliver the results of the computation tothe output driver system under the program control.

The memory includes both a random access memory (RAM) and a read-onlymemory (ROM), the latter storing the information that comprises thecontrol logic. The result of the computations carried out on the inputdata is stored in RAM where it can be addressed, fetched, erased,rewritten or changed, depending upon the operating conditions of thevehicle. The CPU portions that perform the computations comprise anintegrated circuit distinct from the microprocessor chip that comprisesthe memory portions. The memory and the CPU computation portions areconnected by internal bus and interface circuitry.

Data may be read from a first memory location as the processor seeks aninstruction from the memory. The fetched data is then fed into a dataregister or storage area and then to an instruction decoder. When aninstruction is to be carried out, the data that is fetched istransferred to an arithmetic logic unit. Then, sequentially pursuant toinstructions in the instruction decoder, other data may be fetched frommemory and fed into the data registers. The data may be a shift timedelay value, for example, and may be stored in an accumulator until itis addressed during shift sequencing of the processor.

Also sequentially, the data in the accumulator may be transferred to thedata register and then fed back into memory and stored in the nextmemory location within the random access memory (RAM) where it may beaddressed during the next background loop.

The data that is stored in ROM memory may, for example, be shiftschedule information or functions in which two variables, such asthrottle position and vehicle speed, are related one to the other inaccordance with a shift function. The data also may be in the form ofinformation in a table containing three variables or data such as atimer value and values for the other two pieces of data or variables.

The control strategy for the transmission is divided into severalroutines and control modules which are executed sequentially in knownfashion during each background pass. The strategy for each module isexecuted furthermore in sequential fashion, just as the modulesthemselves are executed in sequential fashion. The various dataregisters are initialized as input data from the previously mentionedsensors are introduced into the input signal conditioning portion of theprocessor. The information that results from the inputting of the sensordata, together with information that is stored in memory and learnedfrom a previous background pass, is used to carry out the controlfunctions of the shift solenoid valves, the throttle pressure solenoidvalve, and the bypass clutch solenoid valve. The modules and sub-modulesare executed in sequence in each background loop. Each module or logicportion is independent of the others and performs a specific function.They are executed as they are separately addressed by the processorpointer. These functions occur after the input signals are received bythe input gates and the signal conditioning portions of the processorand after the input signal conditioning has occurred.

The processor logic is established by independent logic modules that arecalibrated for special purposes as will be described subsequently. Forexample, special logic is required to effect maximum smoothness duringstart-up from a standing start as the clutch and brake servos areactuated for acceleration from a standing start or, for example, duringa transition from a forward driving condition to a reverse drivingcondition which requires release of a forward torque delivery clutch orbrake and the engagement of a reverse clutch or brake. Another modulewill establish a desired circuit pressure when the vehicle is coastingor when the vehicle is operated in a power-off mode and a shift iscalled for by the control logic. The circuit pressure is dependent uponengine torque as represented by throttle position, but the capacity ofthe clutches and brakes cannot be made solely dependent upon thethrottle position since, under certain operating conditions such ashigh-speed coasting or power-on or power-off ratio changes, an augmentedcircuit pressure is required to avoid excessive friction elementslipping.

Another module will determine the normal TV (EPC) pressure that isrequired when a special condition such as start-up, coast boost orpower-off shift is not present. The normal TV pressure (EPC) calculationdeveloped by this module will take into account the inertia torqueresulting from a change in engine speed during a shift. It also takesinto account the dynamic conditions which require modification of TVpressure (EPC) resulting from variations in engine speed during a shift.

Thus, the normal calculation for determining the torque proportionalpressure, which herein is referred to as throttle valve pressure or TVpressure, will consist of both a static portion and a dynamic portion.The static TV portion is determined by the torque input to thetransmission and compensates for various factors that affect the nettorque such as the ambient pressure, accessory loads, etc. The dynamicportion of the TV calculation takes into account the inertia torque dueto changes in rotary velocity of the engine and other rotating portionsof the driveline during a shift. The dynamic portion is added to thestatic portion to obtain a total value that is representative of torqueat any given instant. This corrected value is necessary in order toachieve optimum shift feel and to eliminate harshness that might beperceived by the vehicle operator.

The static TV portion is calibrated by this normal TV calculation modulefor each gear. This requires a calculation of the effective torque,which is obtained by reading from a table stored in computer memory avalue that is a function of engine speed and load, the latter being ameasure of air charge. From that value, accessory losses are subtracted.The result of that subtraction is multiplied by the torque ratio of thetorque converter which exists for each torque converter speed ratio, thespeed ratio being a ratio of the turbine speed to the engine speed.

The calculation in this module will determine, for each gear, the slopeof the functional relationship between static TV pressure and the statictorque. During a shift, the dynamic torque is added to the input torqueprior to the multiplication by the converter torque ratio. A differentshifting slope constant is used for each gear. The dynamic TV pressureportion is added to the static portion of the throttle valve pressureduring a shift.

Part of the pressure control strategy includes modulation of the clutchpressure during an upshift to improve smoothness during the shift. Thisinvolves using a converter turbine speed derivative to trigger apressure drop so that the capacities of the clutch and brake servos aretemporarily reduced to a level that will contribute to maximumsmoothness but which will be sufficient to maintain adequate torquetransfer capacity during the shift interval.

The ability of the clutches and brakes to transmit torque depends, ofcourse, on the level of the pressure maintained in the control circuitby the main pressure regulator. This control is unlike TV pressurecontrols of conventional transmissions which rely upon mechanicalthrottle valve linkages to maintain a desired throttle valve pressure ora vacuum diaphragm which is actuated by engine intake manifold pressure.The TV control in the present design is achieved by a variable forcesolenoid valve, shown in FIG. 25, that responds to a signal developed bythe electronic microprocessor. Electronic TV strategy for the processorincludes the step of looking up engine torque from a table and varyingappropriately the signal delivered to the variable force solenoid toadjust the torque transmitting capacity of the transmission.

The static capacity, which as mentioned previously is one of thecomponents of the throttle valve TV pressure, is equal to the throttlevalve (EPC) pressure required to hold the weakest friction elementbecause of engine combustion torque (torque net) and inertia torqueduring a shift. When a shift is not occurring, the capacity is equalmerely to static capacity. The sum of the two torque components ismultiplied by the torque converter torque ratio to obtain the totaltorque capacity requirements. Adjustments to the static torque are madeby determining the dynamic TV, as explained previously, in order toobtain the desired shift feel, to make adjustments for rapid "tip-inshifts" and to counteract for the lag time required for the variableforce TV solenoid to respond.

The throttle valve pressure may be modified during a shift by retardingthe engine spark at the proper instant. Throttle valve pressure isdetermined by a so-called TV-Guide throttle valve pressure module onceevery background loop, which may be as long as 100 milliseconds.However, a throttle valve pressure modification may occur at thebeginning of a shift without waiting for the full 100 millisecondbackground loop to be completed. The so-called TV-GUIDE module in thesecircumstances will be executed as the next sequence. At initiation of ashift, an engine ignition spark retard may be called for, but thisoccurs during a 100 microsecond foreground loop or repeater loop asdistinct from the 100 millisecond background loop. A delay between thisspark retard and the execution of the throttle valve module should beavoided so that these events occur at the same time whereby the TVpressure may be allowed to drop before a rise of inertia torque, whichmight be felt by the vehicle operator.

When an upshift is commanded, the torque transmitted is transferred fromone friction element to another. During the torque transfer, the driventorque will be increased during the so-called inertia phase following adecrease of driven torque during the preceding torque phase of theshift. Provision is made for retarding the spark at the beginning of theinertia phase, and the throttle valve pressure module routine will beexecuted before any other module is addressed in the background loop.The normal sequencing thus is interrupted in order to permit the TVpressure reduction to occur immediately.

In the case of the shift control modules, the four main modules are thePRNDL Based Desired Gear Determination module, the PRNDL Based CommandedGear Determination module, the Load Shift in Progress Timer module, andthe Determine Shift Solenoid States module.

The PRNDL Based Desired Gear Determination module is divided into threesub-modules called the GR₋₋ DS₋₋ PRNDL=3 or 4 module, GR₋₋ SEQ₋₋ PNTRcalculation sub-module, and the Delay Verify Shift sub-module.

The second main module, namely, the PRNDL Based Command GearDetermination module, has four sub-modules which are identified as theGR₋₋ CM,PRNDL=1 Logic sub-module, the GR₋₋ CM,PRNDL=2 Logic sub-module,GR₋₋ CM,PRNDL=3 or 4 Upshift Logic sub-module, and the GR₋₋ CM,PRNDL=3or 4 Downshift Logic sub-module.

The PRNDL Based Desired Gear Determination module determines the gearthat the transmission should assume to satisfy a given set of steadystate operating conditions. For example, it will choose the appropriategear or the desired gear for certain throttle openings and vehiclespeeds and for a given road load, but it will change if any of theprevailing conditions should change. If the selector and lever PRNDLposition is manual 1, the desired gear is set to 1.

If the manual lever PRNDL is in neutral, the desired gear is set to 1.

If the manual lever is placed in the overdrive position or the driveposition, then sub-module GR₋₋ DS₋₋ PRNDL=3 or 4, the GR₋₋ SEQ₋₋ PNTRcalculation sub-module, and the Delay Verify Shift sub-module areexecuted. If a vehicle shift is taking place at this time, the sequencewill be interrupted.

The sub-module GR₋₋ DS₋₋ PRNDL=3 or 4 has a sub-routine that selects thedesired gear using stored information based upon vehicle speed andthrottle position. The desired gear is set to the next higher gear ifthe engine speed is greater than the sea level wide-open throttle engineupshift speed for the current gear.

The next module is the Delay Verify Shift Module. If a change occurs inthe desired gear, this module will delay a change to that gear until anew desired gear has been present a sufficient length of time to verifythat a shift is appropriate. For example, if the shift is the result ofa transient condition that begins and ends during a time less than thetime required for a delay timer to expire, a change in desired gear willnot occur. Further, the delay will permit the new desired gear to bedelayed by a time that is determined by a sub-module called "Load TM₋₋VER₋₋ SFT for Upshifts", which determines the delay time required toverify that an upshift should be made. If a fast "backout" rate isdetected during this routine, for example, the time required to verify atip-out upshift is used. Otherwise, the time required to verify a normalupshift is used.

Next, the processor addresses the module called GR₋₋ SEQ₋₋ PNTR whichcarries out a gear sequence calculation. This module determines howmulti-step shifts will be sequenced based upon a series of calibrationparameters. Each step of every multi-step shift has a calibrationparameter that is unique to it. For example, if the current gear is 1and the desired gear is 4, a 1-4 upshift is called for. There are threecalibration parameters; namely, GR₋₋ SEQ₋₋ 141, GR₋₋ SEQ₋₋ 142, and GR₋₋SEQ₋₋ 143 to determine which gear to command during each step of theshift. This will result in a 1-2-3-4 shift, a 1-2-4 shift, a 1-3-4shift, or a 1-4 shift.

The next main routine is carried out by the PRNDL Based Commanded GearDetermination module which determines which gear should be commanded. Ifthe range selector (PRNDL) is in the manual low position, the GR₋₋ CM₋₋PRNDL=1 logic is executed. If the PRNDL is in 3 or 4 and an upshift isdesired, the GR₋₋ CM₋₋ PRNDL=3 or 4 upshift control logic is executed.If the manual lever PRNDL is in 3 or 4 and a downshift is desired, GR₋₋CM₋₋ PRNDL=3 or 4 downshift logic is performed. If PRNDL is in neutraland vehicle speed is high, the commanded gear is set to 3, as explainedpreviously. Otherwise, the PRNDL, when in neutral, commands the gear tobe set at 1.

The GR₋₋ CM₋₋ PRNDL=1 module (Gear Command) determines the commandedgear when the range selector (PRNDL) is in manual low. Downshifts arescheduled based upon vehicle speed, and no upshifts are allowed. Whenthe manual valve is moved to manual-low from fourth gear or third gear,the transmission immediately shifts to second gear with engine braking.The transmission remains in second gear until a timer runs down andvehicle speed falls below some calibrated speed. The transmission thenshifts to first with no permitted while the manual valve remains inmanual-low.

The next sub-module that is executed in this routine is the GR₋₋ CM₋₋PRNDL=3 or 4 upshift module (Gear Command). This sub-module determinesthe commanded gear when the range selector (PRNDL) is in 3 or 4 and anupshift is desired. If a single step upshift is desired, the commandedgear is set to the desired gear. If a multi-step upshift is desired, thecommanded gear is set to the first step of the multiple ratio shift, asdetermined by the calculation of the GR₋₋ SEQ₋₋ PNTR calculationsequence. Again, shifts from one step to another are delayed by thesequence time determined by the upshift delay logic.

The upshift delay logic is a routine executed by the sub-module thatdetermines the time that should be spent in each gear before the nextstep of a multiple step upshift can be executed. If an upshift from thesecond gear to the third gear is desired, the sequence time for thatshift must expire before the upshift can be completed. If the currentgear is 3, the sequence timer is set to a new value before the executionof the 3-4 shift can be completed.

Following completion of the logic steps for an upshift, the GR₋₋ CM₋₋PRNDL=3 or 4 Downshift (Gear Command) module will address the downshiftdelay logic and will execute it if it is called for. This moduledetermines the time that must be spent in each gear before the next stepof a multi-step downshift can be executed. In this respect, its functionis similar to the upshift delay logic function described previously. Ifthe current gear is 3, the sequence timer is set to the sequence timebefore the execution of the 3-2 shift. If the current gear is 2, thesequence timer is set to the sequence time before the execution of the2-1 shift.

Each time a new gear is commanded, the Load Shift In Progress Timermodule loads a unique time into the shift in progress timer. Calibrationconstants are provided for power-on upshifts, power-off upshifts,power-on downshifts, power-off manual downshifts, and power-offautomatic downshifts. This timer must count down to zero before thePRNDL Based Desired Gear Determination module will begin its executionfor a new desired gear while the range selector (PRNDL) is in 3 or 4.

Shift Logic

A description of the shift logic for normal upshifts now will be made.That will be followed by a general description of the shift logic fornormal downshifts. These are examples of how the control logic willreact to various changes in the operating conditions that areencountered during a typical driving maneuver with the driver-controlledmanual range selector valve (PRNDL) in 3 or 4.

A check is made by the processor for an appropriate vehicle speedcontrolled upshift. All schedules for upshifts into the higher ratiosare checked and the desired gear is set to the highest gear that isallowed by the shift schedules. If the vehicle speed is higher than theupshift function for throttle position corrected for altitude and othervariables, then an upshift is commanded.

The new desired gear is not allowed to pass through to the commandedgear module until the upshift verify timer has run down. This delay isachieved by the previously mentioned logic called "Delay/Verify ShiftLogic". When the new desired gear is passed through to the PRNDL BasedCommanded Gear Determination module, the GR₋₋ CM₋₋ PRNDL 32 3 or 4 logicis executed. In that case, if this is the first execution of the logic,the commanded gear is set to the desired gear. If this is a single stepshift, that logic then is finished. If it is a multiple-step shift, anew sequence timer is loaded to effect a shift delay time for that partof the multiple-step shift. Similar delay times are introduced for eachother step of the multiple-step shift. When all of the delays arecompleted, the commanded gear is set to the gear for the final step ofthe shift.

If the desired gear is one greater than the current gear, a single-stepshift is desired. If it is more than one greater than the current gear,a multiple-step shift is desired. The previously mentioned calculationcalled "GR₋₋ SEQ₋₋ PNTR Calculation" will determine the first step ofthe shift. If the shift is a tip-out shift, which results from a quickrelaxation of the engine throttle, the new desired gear value is notallowed to pass through to the commanded gear module until the tip-outupshift verification time has passed. Thus, a timer will preventcontinuation of the logic until the tip-out upshift verification timehas elapsed.

If the upshift desired is not a tip-out upshift, the new desired gear isnot passed to the commanded gear module until the upshift verificationtimer for that upshift has expired. When that occurs and the shift delayis over, the commanded gear is set to the gear for the second step of amultiple-step shift. If still another step is required to complete theshift sequence, the shift sequence timer for the next step is loaded anda second delay, independent of the first, delays the next step. Whenthat shift sequence time is expired, the commanded gear is set to thegear for the final step of the multiple-step shift.

During normal downshifts, if there is no shift in progress, a check ismade to see if there is a new desired gear. The processor checks allschedules for downshifts into lower gears and sets the desired gear tothe lowest gear allowed by the shift schedules. As in the case ofupshifts, if the desired gear is one less than the current gear, thecommanded gear is set to that desired gear. If the desired gear is morethan one less than the current gear, the GR₋₋ SEQ₋₋ PNTR calculationdetermines the first step of the shift. Each step of the shift has anindependent calibration constant.

The Control Valve Circuit

Shown in FIGS. 4A and 4B is a valve circuit illustrated in schematicform. These views are enlargements of FIG. 4. The valve elements in thevalve circuit of FIG. 4 are in the positions they assume when thetransmission is in park, the engine throttle is closed and the converterclutch is off.

The circuit of FIG. 4, 4A and 4B includes a main regulator valve andbooster shown at 124. This valve regulates line pressure fromengine-driven pump 126. This pump was described generally in FIG. 1. Itincludes Gerotor pump elements 38 as mentioned previously.

The main regulator valve 124 determines the line pressure in linepressure passage 128. It responds to an electronic pressure controloutput delivered to the main regulator 124 through line 130. Thispressure is developed by the electronic pressure control solenoid valve114 as will be described subsequently.

As seen in FIG. 4A, main regulator valve 124 comprises a valve spool 132having spaced lands 134, 136, 138 and 140. These lands register withinternal lands when in the main regulator valve bore as shown. A boostervalve spool 142 is situated in a sleeve in the main regulator valvebore. The booster valve sleeve provides a seat for main regulator valvespring 144, which tends to urge the main regulator valve spool 132 in anupward direction against the opposing force of line pressure acting onthe upper end of land 140 as shown at 146.

The lower end of land 134 is subjected to the pressure in the electronicpressure control line 130, thus producing a force that assists thespring 144.

Booster valve spool 142 has a land 148 that is subjected to pressuredelivered to the booster valve spool through reverse line pressure line150, which in turn communicates with reverse line pressure passage 152.Passage 152 extends to the manual valve generally indicated at 154 inFIG. 4B. The manual valve will establish pressure distribution to line152 when the manual valve resumes the reverse drive position R.

The main regulator valve distributes pressure to the converter andlubrication circuits through passage 156 as seen in FIG. 4A. Valve land136 establishes controlled communication between line pressure passage128 and passage 156 in response to the pressure and spring forces actingon the regulator valve spool. The magnitude of the pressure in passage128 is determined by the regulating action of land 134 which establishescontrolled communication between passage 128 and return passage 158leading to the intake side of the pump 126.

The main regulator valve 124 regulates line pressure by exhausting aproper amount of oil from the pump outlet line 128. When the valve spool132 is in its uppermost position, the pressure will rise until it islimited by either leakage or available power input to the pump. At theother extreme position of the regulator valve spool, the valve is fullyopened and all pump flow is exhausted. The valve spool 132 willautomatically position itself between these two extremes, thus creatinga variable restriction by exhausting the proper amount of oil into line158 to maintain the desired pressure. A pressure increase occurs as thevalve moves upwardly in response to a pressure signal from theelectronic pressure control solenoid.

The regulator valve regulates the position of the valve spool bybalancing electronic pressure control output pressure in line 130 andthe force of the spring against the line pressure force acting at theupper end of the valve on land end 146. If the line pressure is lessthan the desired amount or if the electronic pressure control pressurein line 130 increases, the force balance is upset and the pressuresignal from the electronic pressure control solenoid and the springforce will move the valve upwardly, thus reducing the exhaust flow andincreasing the line pressure. The converse action will occur if the linepressure is too high or if the electronic pressure control signal inline 130 is decreased.

The regulator valve will respond continuously to changes in pump flow tocause a readjustment of the position of the valve to maintain a presetbalance. A decrease in pump flow, or an increased demand for pump flow,during clutch application for example, will cause the valve to closeslightly and exhaust less oil in order to maintain the desired pressure.

When the engine is started, the main regulator valve opens to a pointwhere the flow is delivered first to the converter circuit through line156. This flow goes through the converter limit and bypass controlvalves into the converter and lubrication circuits. These circuits arecharged quickly and the regulator valve opens further until the flow isexhausted into the passage 158 which recirculates the oil back to thepump intake. If there is not sufficient pump flow to satisfy the flowdemands for both line pressure and converter pressure, the mainregulator valve will prioritize the pump flow to maintain the scheduledline pressure.

The line pressure acting on land end 146 is routed through orifice 160,which dampens rapid valve movements which would cause pressureoscillations.

The main regulator valve spring 144 prevents line pressure from beingtoo low at low output pressures of the electronic pressure controlsolenoid. If the electronic pressure control signal in line 130 is abovea predetermined valve, such as 19 psi, the booster valve is shifteddownwardly into a sleeve 162 so that it does not contact the mainregulator valve spool 132. When the output of the electronic pressurecontrol solenoid is less than 19 psi, for example, the booster valvewill contact the main regulator valve spool.

During reverse drive, passage 150 is pressurized as mentioned earlier.The force of that pressure is added to the force of the booster valvespring.

Oil from the torque converter passes through line 164 to drainback valve166, which comprises a movable single-diameter valve element 168. Thisvalve element is urged downwardly under its own weight in a valvechamber to control the effective flow through drain back orifice 170.The output flow from the drainback valve passes through line 172 to thetransmission oil cooler 174. When the engine is turned off, the weightof the valve seats to block the oil from draining out of the converterthrough line 164.

The converter pressure limit valve shown at 176 limits the pressure ofthe converter circuit by controlling the amount of oil delivered to theconverter through line 156 by the main regulator valve. At lowerpressures, the valve is held in an uppermost position by valve spring178 acting on converter pressure regulator valve element 180. Convertercharge pressure then flows freely to converter feed line 182 through thespace between adjacent lands 184 and 186. If the converter chargepressure in line 182 rises above a predetermined value, the valve spool180 will be forced outwardly by reason of the pressure acting betweenlands 186 and 188. The magnitude of the pressure in passage 182 thenwill be regulated at a value determined by the spring force.

The manual valve 154, seen in FIG. 4B, receives line pressure throughpassage 190, which communicates with line pressure passage 128 describedpreviously. The manual valve 154 includes a driver-controlled valvespool 192 having spaced lands 194, 196 and 198. The valve spool hasseveral positions that may be selected by the operator. These positionsare determined by detents and are identified in FIGS. 4 and 4B by thesymbols P, R, N, OD, D and 1, which respectively identify the Parkposition, the Reverse position, the Neutral position, the Overdrivedrive range position, the three-speed ratio D range position, and themanual low range position. The valve is actuated by a driver-operatedmanual shift valve linkage. When the valve spool 192 is positioned asshown in FIG. 4B, the valve circuit is conditioned for park.

Line pressure from passage 190 is transferred through the space betweenlands 196 and 198 to the reverse drive circuit through passage 200, tothe overdrive circuit through passage 202, passage 204 and passage 206,to the three-speed ratio drive range circuit through passage 204 and 206and to the manual low drive range circuit through passage 206.

Passage 208 is pressurized by the manual valve when it assumes the parkposition, the D position, the reverse position or the "1" position.

Line 210 is an exhaust passage located between the lands 196 and 194.

The bypass clutch control valve, which is generally indicated at 212,comprises a movable valve spool 214 with equal diameter spaced valvelands 216, 218 and 220. A larger diameter valve land 222 provides adifferential area on which bypass clutch pressure in passage 224 acts.

The bypass clutch control valve controls the operation of the converterbypass clutch. When the clutch is not engaged, pressure from passage182, described previously, is distributed through the bypass clutchcontrol valve to the converter supply line 226 and to the bypass clutchsupply line 224. The modulated converter clutch control solenoid is offat that time, as will be explained subsequently.

When the bypass clutch control valve spool 214 is in an upward position,the clutch piston plate, shown at 42 and 44 in FIG. 1, is held out ofengagement with respect to the adjacent friction surface of the impellerhousing 26. All of the torque delivered to the gearing from the enginethen is delivered hydrokinetically.

The converter clutch is engaged, upon actuation of the modulatedconverter clutch control solenoid, as the bypass clutch control valveelement 214 moves downwardly under the influence of modulated converterclutch solenoid pressure in passage 228, which acts on the upper end ofthe land 216. The valve spool 214 then regulates the pressure in passage224, which extends to the bypass clutch, the magnitude of the pressurein passage 224 being determined by the magnitude of the modulatedconverter clutch solenoid pressure in passage 228 and by the pressure inpassage 182.

The microprocessor monitors continuously the torque converter slip andcontinuously adjusts the magnitude of the modulated converter clutchsolenoid duty cycle to achieve the desired output pressure in passage228 so that the slip is controlled to a predetermined value as will beexplained subsequently.

Line 164 returns oil from the torque converter to the cooler through thedrain back valve as explained previously. When the bypass clutch is notengaged, oil flows freely through the bypass valve from the line 164 tothe cooler circuit. When the valve is engaged, the oil flow isrestricted by orifice 230.

When the modulated converter clutch solenoid is deenergized, its outputpressure in line 228 is zero and the bypass clutch control valve ismoved upwardly by valve spring 232 and by pressure from passage 182acting on the end of valve plunger 234 located at the lower end of thebypass clutch control valve and directly engaging the bypass clutchcontrol element. The pressure in passage 182, as explained previously,is limited by the converter pressure limit valve.

When the modulated converter clutch solenoid is off, oil flows freelyfrom passage 182 to passage 224 and through flow restricting orifice 236to the converter impeller circuit line 226. This is done to restrict theflow in passage 226 so that the converter clutch will not drift on whenclutch application is not desired.

The converter turbine circuit, which includes lines 164, carries oilfrom the converter to the cooler and the lubrication circuit asexplained previously. During open converter operation, the flow passesunrestricted through the bypass clutch control valve and the drainbackvalve.

When the modulated converter clutch solenoid is actuated, a modulatedpressure is developed in passage 228. A pressure in excess of 6 psi in apreferred embodiment is required to overcome the force of the valvespring 232. When the bypass clutch control valve is in its modulatingposition, flow from line 182 to the bypass clutch supply passage 224 isbypassed and fluid flow is unrestricted into the converter impellercircuit through line 226 to maintain full pressure on the rear of theconverter clutch piston shown at 42 and 44.

Converter bypass pressure acts on the opposite side of the piston, asshown at 42 and 44, and is regulated by the bypass clutch control valveas a function of the pressure in the passage 228 and by the force of thespring 232 and the force of pressure from passages 182 acting on valveplunger 234.

The microprocessor, during modulation of the pressure in passage 224,calculates turbine speed from the output shaft speed and the gear ratioand compares that value with the engine speed to determine converterslip. The pressure in passage 228 is continuously adjusted to maintainthe desired slip. To reduce the slip, the microprocessor commands ahigher value of pressure in passage 228 which reduces the pressure inpassage 224, thus increasing the clutch torque capacity. To eliminateslip, the microprocessor can command a maximum pressure in passage 228.The solenoid acts with a 100% duty cycle at that time, thus moving thebypass clutch control valve downwardly to maintain the pressure inpassage 224 at a zero value since it is brought into communication withexhaust port 238.

The shift solenoids SS1 and SS2 are illustrated in FIG. 4B in blockdiagram form. They are identical in form, as will be explainedsubsequently with reference to FIGS. 27 and 27A.

Solenoid SS1 is fed with pressure from solenoid supply passage 240through a flow control orifice 242. Solenoid SS2 is supplied withsolenoid supply pressure from passage 240 through orifice 244. Thesolenoid supply line for the solenoid SS2 on the downstream side of theorifice 244 is shown at 246. The corresponding supply line for solenoidSS1 is shown at 248. When the shift solenoid is deenergized, the flowthrough the orifice is exhausted and the output pressure for solenoid S1or S2 is zero. When the solenoid is energized, however, the solenoidblocks the flow and the solenoid output pressure rises until it equalssupply pressure.

Referring next to FIGS. 27 and 27A, solenoid SS1, which is identified inFIG. 3A by numeral 120, includes electrical windings 250 and an armature252. (Since solenoid SS2 is similar to solenoid SS1, FIGS. 27 and 27Amay be considered to apply also to solenoid SS2.) The armature engagesball valve element 254, which acts on and registers with an orificeformed in valve seat 256. Solenoid spring 258 urges normally thearmature 252 in an upward direction. When the solenoid is off, valve 254moves off its seat thereby exhausting pressure from supply line 248 toan exhaust port 260.

When the solenoid 120 is energized, the armature 252 moves to theposition shown in FIG. 27A, thereby seating ball valve element 254against its valve seat and interrupting communication between passage248 and exhaust port 260. This causes a shift signal that will beutilized by the shift valve elements to be described subsequently.

The pressure supplied to the shift solenoids through passage 240 isobtained from solenoid pressure regulator valve 260 shown in FIG. 3B.Pressure is distributed to the valve 260 through supply line 262 whichis pressurized by the manual valve when the manual valve is in theoverdrive position, the D position, or the 1 position.

The valve 260 includes a regulator valve element 264, which is urged ina left-hand direction by valve spring 266. Valve element 264 haspressure regulating valve lands 268 and 270 which control communicationbetween passage 262 and passage 240. A feedback branch passage 272 actson the left-hand end of the solenoid pressure regulator valve spool 264.Thus, the pressure in passage 240 is functionally related to the springforce of valve spring 266. The pressure in passage 240 is a constantvalue in all forward drive ranges. Passage 240 also supplies a modulatedconverter clutch solenoid 116 with supply pressure as will be explainedsubsequently.

The 1-2 shift valve is identified in FIG. 4B by reference numeral 274.It includes a shiftable valve element 276 which is located in directalignment with valve element 278 for the 2-3 shift valve generallyindicated at 280. The 1-2 shift valve controls ratio changes betweenfirst and second gears. When the first solenoid SS1 is energized,solenoid pressure is developed in line 248, as explained previously.This pressure is distributed to the upper end of land 282 therebyshifting the valve spool 276 downwardly against the opposing force ofvalve spring 284. That spring is effective also on valve land 278 forthe 2-3 shift valve 280.

The 1-2 shift valve includes also lands 286 and 288.

Passage 206 which is pressurized during operation in the overdriverange, the D range and the manual low range is blocked by land 282during first ratio operation. All of the output circuits are exhaustedthrough exhaust port 290. Thus, line 294 which extends to theintermediate clutch and the overdrive servo supply, is exhausted.

Passage 292 which is pressurized by the manual valve in the park,manual-low range and reverse drive positions of the manual valvedistributes pressure through orifice 296, through the space betweenlands 286 and 288 and through check valve 298 to the low and reverseservo 300. When the 1-2 shift is initiated by the processor, shiftsolenoid SS1 is deactivated and the 1-2 shift valve is moved by thevalve spring upwardly. Pressure from passage 206 then is delivereddirectly to the line 294 which distributes pressure through passage 302to the intermediate clutch 304. Communication between passage 294 andpassage 302 is established by orifice 306. Pressure is delivered also tothe 1-2 accumulator through passage 308. Pressure also is delivered atthat time from passage 294 through flow control orifice 310 to passage312 leading to the overdrive servo regulator valve generally identifiedby reference numeral 314.

The 2-3 shift valve 280 controls shifting between the second ratio andthe third ratio. During operation in the second ratio, shift solenoidSS1 and shift solenoid SS2 are off and the spring holds the 2-3 shiftvalve element 278 in a downward position. It also holds the 1-2 shiftvalve element 276 in an upward position. Direct clutch feed line 316during second ratio operation is exhausted through exhaust port 318, thelatter communicating with line 316 through the space between lands 320and 322 on the 2-3 shift valve spool 278. When the manual valve is inthe overdrive range or the D range, oil from the manual valve flowsthrough the 2-3 shift valve from passage 324 to the release side of theoverdrive servo 326. Communication between the 2-3 shift valve and therelease side of the overdrive servo is obtained from overdrive releasefeed passage 328. This assures that the overdrive servo will remain offeven when the apply side of the overdrive servo is pressurized. Inmanual-low (1) drive range, the overdrive servo release pressure isexhausted through the manual valve and through passage 204, passage 324and passage 328 so that when the overdrive servo apply chamber ispressurized, the overdrive servo will be activated.

A ratio change to the third gear is obtained when the processor turns onshift solenoid 2 while shift solenoid 1 remains off. Shift solenoidpressure then is distributed to the lower end of land 320 of the 2-3shift valve and moves the 2-3 shift valve upward. Pressurized oil fromthe manual valve passes through passage 324 to passage 316, which leadsto the 2-3 backout valve generally identified by reference numeral 330.That pressurized fluid then passes directly through the 2-3 backoutvalve to the direct clutch 332 through passage 334. Pressurized fluid inthe direct clutch 332 is transferred to the 2-3 accumulator throughpassage 336. The 2-3 accumulator is identified generally by referencenumeral 338. The pressure then is distributed to the overdrive servorelease chamber through passage 328 and to the overdrive release circuitfrom passage 340 instead of being connected to the passage 324 leadingto the manual valve. This is done to prepare for a 3-4 upshift.

A 2-3 capacity modulator valve is identified generally by referencenumeral 342. This valve, during a 2-3 upshift, regulates the pressure inpassage 344 extending to the upper side of the piston 346 of the 2-3accumulator 338. This then controls the direct clutch pressure on thebottom side of the accumulator piston 346. The pressure in passage 344opposes the force of accumulator valve spring 350.

The 2-3 capacity modulator valve 342 is situated in a common valve boreand acts in cooperation with the orifice control valve 352, whichcomprises a valve spool 354 that is urged downwardly by valve spring 356and the direct clutch pressure in passage 336 during a 2-3 upshift.

The direct clutch pressure in the chamber 348 is determined by theregulated pressure at the top of the 2-3 accumulator piston 346. Duringa 2-3 upshift, pressurized fluid from the 2-3 shift valve flows throughorifice 358 or orifice 360, depending upon the position of the 2-3backout valve to the direct clutch 332. This strokes the 2-3 accumulatorpiston 346 in an upward direction. Prior to the 2-3 shift, theaccumulator piston assumes the downward position under the influence ofthe forward clutch pressure on top of the piston in the first and secondratio. As the 2-3 accumulator piston 346 strokes, the force balance onthe piston control affects clutch pressure. Direct clutch pressure addedto the 2-3 accumulator spring force moves the accumulator piston 346upward, thereby forcing oil out of the accumulator through line 344.This causes ball check valve 362 to close. This forces oil to flowthrough the 2-3 capacity modulator valve, which regulates theaccumulator back pressure to a higher value than the line pressure.

The 2-3 capacity modulator valve has a valve spring 364 which urges the2-3 capacity modulator valve spool 366 downward thereby providingcontrolled communication between passage 344 and passage 368, whichcommunicates with the 2-3 capacity modulator valve element 366 and withthe forward clutch feed line 370.

The 2-3 backout valve 330 has a multiple land valve spool 372. Theoutput pressure of the electronic pressure control solenoid valve 114 isdistributed through line 374 to the left-hand end of valve land 376 ofthe backout valve 330. The force developed by the electronic pressurecontrol solenoid pressure in passage 374 is opposed by backout valvespring 378.

During a 2-3 shift, the 2-3 backout valve feeds pressurized oil to thedirect clutch 332 through either orifice 358 or orifice 360. Normally,orifice 358 is the feed orifice during part throttle operation andorifice 360 is the feed orifice during closed throttle operation.However, since the line pressure established by the EPC control solenoid114, which is controlled by the processor 92, may be affected by factorsin addition to throttle pressure, part throttle 2-3 shifts may useorifice 360 in some cases and orifice 358 in other cases.

The 2-3 backout valve, in combination with the orifice control valve352, controls the apply rate of the forward clutch and the overdriveservo release. During a closed throttle forward engagement of theforward clutch, the forward clutch and the 2-3 accumulator arepressurized through the orifice 380. During power-on engagements, theorifice 380 is bypassed and the feed rate to the forward clutch isunrestricted. Thus, oil will be transferred through passage 382,bypassing the orifice 380, and through the space between backout valvelands 384 and 386 to the forward clutch.

During a 4-3 coasting downshift, the forward clutch is applied slowly asapply pressure passes through control orifice 388 located in the forwardclutch feed line 370. Simultaneously, the flow to the overdrive servorelease through passage 328, which is fed through the 2-3 shift valvefrom passages 390 and 340, is unrestricted and the release occursquickly.

During a power-on 4-3 downshift when the backout valve element 372 isshifted in a right-hand direction, the flow to the forward clutch and tothe overdrive servo release are combined and fed through a commonorifice 392. The feed line 328 for the overdrive servo releasecommunicates with passage 394. Passage 394 communicates with line 370,which is the feed line for the forward clutch.

It is the function of the orifice control valve 352 to control theapplication rate for the forward clutch passage 368 extending from the3-4 shift valve 394 holds the orifice control valve spool in an upwardposition during first and second gear ratio operation. Further, thepressure in passage 368 is transferred through the orifice control valve352 to passage 396, which is the feed passage for the forward clutch infirst and second gear.

The left end of land 376 of the 2-3 back-out valve 230 is subjected toelectronic pressure control solenoid pressure in passage 374. In FIG.24A is shown a plot of the electronic pressure control solenoid pressureon a vertical axis and throttle position on a horizontal axis. Forthrottle settings between zero and an initial throttle position shown at670, the throttle valve pressure is at a fixed value greater than zero,but less than the maximum shown at 672. This is shown by a horizontalline 674. For a value of throttle valve pressure equal to that indicatedat 674, the 2-3 back-out valve will not be shifted to the right sincethe pressure force thus developed is insufficient to overcome the forceof the valve spring 378. For various commands that are issued to effectchange in ratio, a different pressure will be made available to the lefthand of the land 376.

The strategy for accomplishing these back-out valve functions is shownin FIG. 24B. The processor will continuously monitor during eachbackground loop the conditions for establishing the appropriate pressurelevel for the 2-3 back-out valve by inquiring first at step 680 in FIG.24B whether the control system is making a change that requiresengagement of the forward clutch. If the answer is yes and the shift inplace is a power-on shift, a maximum value for the minimum TV pressureis determined at point 676 shown in FIG. 24A. Further, a minimum value678 is determined if the engagement in place occurs during a power-offshift. If the inquiry at step 680 is negative, the routine will proceedto step 682 where an inquiry is made to determine whether a first gearis being commanded. If that is case and if the shift is a power-onshift, a different minimum value for point 682 is determined. If it is apower-off shift, a fixed TV value is used. If the inquiry at step 682 isnegative, an inquiry then is made at step 684 to determine whether asecond gear is being commanded. If that is the case, a still differentminimum value or TV pressure is calculated for point 676 while aconstant TV value is maintained for power-off shifts.

If the inquiry at step 684 is negative, the routine will proceed to makean inquiry at step 686 where it is determined whether third ratio isbeing commanded. If that is the case, a still different minimum value orTV pressure is calculated for a point 676. Again, the minimum value fora power-off shift is fixed.

If the fourth ratio is being commanded rather than first, second orthird ratios, and the shift is a power-on shift, a still differentminimum value is calculated for point 676 as shown at step 688. Again,if the shift at 688 is a power-off shift, the fixed value for TV isused.

If during the shift the TV pressure is greater than the value at 672,the greater value that is used is the TV pressure-base on the TVcalculation that is used. This calculation results in the curve shown at690 in FIG. 24A. If, on a power-on shift, the value does not exceed thevalue at 676, the 2-3 back-out valve will not move. If the shift is apower-off shift and the value does not exceed the value at 678, the 2-3back-out valve will not move.

Shown in FIG. 24C is a diagram that summarizes the orifices that are inplace as the 2-3 back-out valve functions. This 2-3 back-out valve logicwill introduce orifice 380 in the fluid flow path leading to the forwardclutch through passage 370 during the engagement logic for the forwardclutch. On the other hand, as indicated in the logic diagram of FIG.24C, the fluid flow path to the overdrive release servo is a free flowpath with no control orifice.

The 4-2 kickdown logic, as indicated in the logic diagram of FIG. 24C,does not result in the introduction of a flow control orifice by the 2-3back-out valve into the feed flow paths for the forward clutch and theoverdrive release side of the overdrive servo.

During a coasting 4-3 downshift, the logic will result in introductionof orifice 388 into the flow path to the forward clutch through passage390. The flow path to the overdrive release side of the overdrive servo,however, is free of a flow restricting orifice as indicated in the chartof FIG. 24C.

During a power-on 4-3 downshift, the 2-3 back-out logic diagram of FIG.24C indicates that the forward clutch is fed through orifice 392 frompassage 398. The fluid flow path to the release side of the overdriveservo is also on the downstream side of the orifice 392. It is apparentfrom FIG. 24C that on a coasting 4-3 downshift, the overdrive releasewill be pressurized to release the overdrive brake much sooner during acoasting downshift than in the case of a torque demand downshift, wherethe overdrive release side of the overdrive servo must be fed with oilthrough the orifice 392.

During fourth ratio engagement of the overdrive servo on a 3-4 shift,both the forward clutch and the overdrive release side of the overdriveservo must be exhausted. In the case of the forward clutch and theoverdrive servo release, the exhaust will occur across orifice 417located at the 3-4 shift valve. The overdrive servo release pressure,however, must be regulated by 3-4 capacity modulator valve 420 and bledthrough gasket bleed orifice 430 and a bleed orifice in the servorelease cavity. Thus, the overdrive release has three flow paths, onepath through the 3-4 capacity modulator valve, one path through theorifice 430 before merging with forward clutch fluid and passing throughorifice 417 and one path directly to exhaust in the servo cavity.

During operation in the third and fourth ratios, the direct clutchpressure, plus the force of valve spring 356 for the orifice controlvalve 352, moves the valve element 354 downward and flow from the 3-4shift valve is transferred through the valve 352 to passage 398, whichis the source of pressure for the forward clutch and the overdrive servorelease during third ratio and fourth ratio operation.

The 3-4 shift valve 394 controls shifting between the third ratio andthe fourth ratio. During operation in the first, second and thirdratios, the 3-4 shift valve is held in a downward position by valvespring 400. Pressure from the shift solenoid 2 is distributed to thedifferential area of lands 402 and 404 on the 3-4 shift valve throughpassage 406. Similarly, output pressure from shift solenoid 1 isdistributed to the lower end of the land 404 through passage 408.Neither the pressure in passage 408 nor the pressure in passage 406 issufficient by itself to move the 3-4 shift valve upwardly against theforce of the spring 400. Line pressure from passage 410, which is thesame as the pressure in passage 262, is transferred through the 394shift valve to the orifice control valve since the two are hydraulicallyconnected through passage 368. Passage 410 then becomes the source ofpressure for the forward clutch. Further, it is a source of pressureduring third ratio and fourth ratio operation for the overdrive servorelease.

A ratio shift to the fourth ratio from the third ratio requires bothshift solenoids SS1 and SS2 to be energized. The pressures acting on thelower end of the land 404 and on the differential area of lands 404 and402 then is sufficient to move the spool valve element 412 for the 3-4shift valve 394 in an upward direction. This conditions the transmissionfor fourth ratio operation. Forward clutch pressure and overdrive servorelease pressure during power-on 3-4 shifts then are exhausted throughcheck valve 414 and through exhaust port 416 in the 3-4 shift valve 394.For closed throttle 4-3 shifts, the servo release pressure does not flowthrough check valve 414. A control orifice 417 is located at the exhaustport 416. The overdrive servo piston then applies as pressure builds upin the servo apply side of the overdrive servo. The rate of pressurebuild up is controlled by orifice 310 in the passage 312.

The 3-4 capacity modulator valve 420 is a simple pressure regulatorcomprising a valve spool 422, which is urged downwardly by valve spring424. It functions to regulate the release of pressure of the overdriveservo during 3-4 upshifts. The overdrive release pressure is distributedto the lower end of the valve element 422 through passage 426, whichextends to the overdrive release passage 328 in third and fourth ratios.Passage 328 and passage 426 communicate with each other through passage340 and the 2-3 shift valve.

As the overdrive servo applies during a 3-4 shift, the flow from therelease side of the servo seats the ball check valve 428 so the releaseoil is forced to flow through the 3-4 capacity modulator valve throughpassage 426.

A small bleed passage is used to bypass the 3-4 capacity modulator valveas shown at 430 to keep the overdrive servo release pressure duringsteady-state operation at zero so as not to compromise the overdriveband static capacity. A second bleed exists in the servo release cavityfor the same reason.

The low servo modulator valve generally identified by reference numeral432 regulates the pressure distributed to the low-and-reverse brake bandservo 300 during first gear manual operation. The manual low drive rangefluid flow from the 1-2 shift valve and through the orifice 296described previously is regulated during first gear operation in themanual low range by the valve 432. Valve 432 is a simple regulator thatmodifies the rate of application of the pressure to the servo 300.

The overdrive servo regulator valve 314 functions to regulate at aconstant apply pressure for the overdrive servo 326 in third gear andsecond gear of the overdrive range in order to smooth the 4-3 power-ondownshifts. Overdrive servo apply pressure acts on the lower end of land436 and balances the spring force. The valve is supplied from the 1-2shift valve through passage 294, which is line pressure in second, thirdand fourth gear ratios.

During fourth ratio operation, pressure from the 3-4 shift valve isdistributed to passage 440 and urges the valve spool 438 downward toprevent it from regulating so that the overdrive servo apply pressureequals line pressure. It functions in this manner also during manual lowoperation so that overdrive servo apply pressure is the same as linepressure whenever the servo apply chamber is pressurized.

The 1-2 accumulator is generally identified by reference numeral 446 andincludes a double-acting piston 448 having an upper pressure chamber 450and a lower pressure chamber 452. Intermediate clutch pressure during a1-2 shift, as explained earlier, is distributed to the chamber 452through passage 308. Line pressure is delivered to the upper chamber 450through line passage 454. This cushions the application of theintermediate clutch. The uppermost chamber of the 1-2 accumulator isexhausted. It could be pressurized, however, if it is desired to furthermodify the intermediate clutch pressure during a 1-2 shift.

The 2-3 accumulator, which was described with reference to theengagement of the forward clutch, functions also to cushion theengagement of the direct clutch during a 2-3 shift. The direct clutch332 communicates with the lower side of the piston 346 during directclutch engagement and the forward clutch 456 communicates with the upperside of the accumulator piston 346.

The control scheme has eight check valves. These include valve 362previously described, valve 414 previously described, check valve 458associated with the 2-3 backout valve, previously described valve 428, avalve 460 associated with the reverse clutch 462, and a check valve 464which actually is a shuttle valve disposed between the low and reverseservo 300 and reverse servo feed line 200 and between the low andreverse servo 300 and the feed line 466 extending to the 1-2 shiftvalve. The seventh check valve is shown at 468 and the eighth checkvalve is shown at 470.

Check valve 362 seats during a 2-3 shift thereby forcing the 2-3accumulator back pressure to flow past the 2-3 capacity modulator valve.During a 3-2 downshift and during a forward clutch engagement, forwardclutch flow unseats the ball and flows to the accumulator.

Check valve 414 seats when the forward clutch is being applied therebyforcing the flow through the appropriate orifice. During a 3-4 shift andduring a disengagement to effect neutral, forward clutch fluid flowsfreely past the valve 414 thus bypassing all of the orifices.

Check valve 458 seats during a 2-3 shift to force the direct clutch flowto go through either orifice 358 or orifice 458, depending upon theposition of the 2-3 backout valve. During a 3-2 downshift direct clutchpressure fluid flows freely past the check valve 458, thus bypassing theorifices.

Check valve 428 seats during a 3-4 shift, thus forcing the overdriveservo release flow to go past the 3-4 capacity modulator valve 420. On a4-3 shift, oil flows freely past the valve 428.

Check valve 460 seats during reverse engagement thus forcing the reversesupply fluid to flow through orifice 472. When the reverse clutch isreleased, fluid from the reverse clutch exhausts freely through thevalve 460.

Shuttle valve 464 applies the low-and-reverse servo from either thereverse circuit or the 1-2 shift valve.

Check valve 468 prevents the distribution of pressure to the converterclutch during first gear operation if for some reason the modulatedconverter clutch solenoid should malfunction during operation in firstgear. The ball is seated in second gear, third gear and fourth gearoperation.

Check valve 470 seats during a 1-2 shift. This forces the intermediateclutch fluid, upon clutch application, to flow through the controlorifice 306. On a 2-1 downshift, a second orifice exhaust path throughthe orifice 474 is added to the exhaust flow path.

FIG. 4A shows the converter circuit and the bypass clutch control whenthe transmission is in the neutral and park ranges at light throttle. Inthis instance, the bypass clutch control valve 212 is in an upwardposition. Regulated pressure then is distributed to the converterimpeller circuit and also to the bypass circuit through passages 226 and224, respectively. Fluid is distributed from the main regulator valvethrough the converter pressure limit valve 176, which assumes an upwardposition. It is transferred then through the bypass clutch control valvedirectly into the passages 226 and 224. Fluid returns from the converterturbine circuit through passage 164 and then through the open drain backvalve to the cooler 174.

FIGS. 5A and 5B show a schematic diagram of the valve elements describedwith reference to FIG. 4, but the valves have been shifted to thepositions that are consistent with reverse-drive, part-throttle,open-converter operation. Thus, the bypass clutch control is in anupward position, the solenoid pressure regulator valve 260 is in aleft-hand position, the 2-3 backout valve is shifted to the right by theelectronic pressure control output pressure in passage 374, the 3-4shift valve element 412 is shifted to a downward position and actuatingpressure is distributed to the low-and-reverse servo 300.

The valve spring 284 moves the 1-2 shift valve element to its upperposition and moves the 2-3 shift valve element to its lower position.

Reverse pressure is distributed to the reverse clutch through passage200 from the manual valve 154.

FIGS. 6A and 6B show the valve elements described with reference to FIG.4, but the valves are moved to the positions that are consistent withneutral, closed-throttle, open-converter operation. Line pressure fromthe manual valve, when the manual valve assumes the neutral position, isdistributed through lines 190 and 156, through the converter pressurelimit valve and through the bypass clutch control valve to the converterimpeller and bypass clutch circuits. The 2-3 shift valve moves down, asdescribed earlier, and the 1-2 shift valve element moves up, asdescribed earlier. Pressure distribution to all of the servos andclutches is interrupted. The 2-3 backout valve and the solenoid pressureregulator valve move to their left-hand positions.

FIGS. 7A and 7B show the valve elements described with reference to FIG.4, but the valve elements have been moved to the positions correspondingto the overdrive, first-gear, closed-throttle, open-converter operation.In this case, the manual valve delivers line pressure through the 2-3shift valve, which is moved in a downward direction, to the feed passage328 for the release side of the overdrive servo. Line pressure isdistributed also to the forward clutch from the 3-4 shift valve, whichassumes its downward position. The 3-4 shift valve is supplied withfluid from the manual valve through passage 410.

FIG. 7B shows in simplified form the valve elements that areparticularly affected by this overdrive operating condition. It is seenfrom FIG. 7B that the upper working chamber of the 2-3 accumulatorpiston is pressurized. The orifice control valve is shifted upward andthe 2-3 capacity modulator valve is shifted downward, therebyestablishing fluid communication between the outlet side of the 3-4shift valve and the forward clutch.

The shift solenoid SS2 is off and the shift solenoid SS1 is on. Thiscauses the 1-2 shift valve to move downward, which in turn forces the2-3 shift valve to move downward because the two valves are connectedmechanically, one with respect to the other.

FIG. 7A shows the position of the bypass clutch control valve duringoperation in the overdrive range with the converter clutch released.

FIGS. 8A and 8B show a valve diagram corresponding to FIG. 4, but thevalves have been moved to their positions consistent with overdrive,second-gear, part-throttle, open-converter operation.

FIG. 8A shows the 1-2 accumulator with both pressure chamberspressurized. Both shift solenoids SS1 and SS2 are deenergized, as shownin FIG. 8B. The manual valve then distributes pressure to the 3-4 shiftvalve, which transfers it to the forward clutch through the orificecontrol valve. The manual valve also distributes pressure to the 1-2shift valve, which distributes it through the orifice 306 to theintermediate clutch.

FIGS. 9A and 9B show a view corresponding to FIG. 4, but the valves havebeen shifted to the positions consistent with overdrive, third gearoperation when the converter is modulating, rather than being open. Inthis instance, the bypass clutch control is moved inward under theinfluence of pressure delivered to it from the modulating converterclutch solenoid. Thus, pressure is distributed through the bypass clutchcontrol valve through bypass clutch feed passage 224. The magnitude ofthe pressure in passage 224 is determined by the duty cycle of themodulated converter clutch solenoid and the pressure in passage 228delivered to the upper land 216 of the bypass clutch control valve.

Pressure from the manual valve is delivered directly to the release sideof the overdrive servo and to the direct clutch through the 2-3 backoutvalve. The 2-3 backout valve is shifted in a right-hand direction by theelectronic pressure control signal in passage 374.

The function of the 2-3 backout valve illustrated in FIG. 4 is betterunderstood by referring to enlarged view of the backout valve as seen inFIG. 9B.

In the condition shown in FIG. 9B, the shift solenoid SS1 is deactivatedand shift solenoid SS2 is activated. A modulated converter clutchsolenoid receives signals of varying pulse width. When the shiftsolenoid 2 is activated, oil pressure is delivered to the 2-3 shiftvalve and the 3-4 shift valve. The 2-3 shift valve shifts upward againstthe opposing force of the valve spring while the 1-2 shift valve remainsupshifted in an upward direction. Oil for the manual valve then flows tothe direct clutch and to the 2-3 accumulator by way of the 2-3 backoutvalve.

The pressure from the shift solenoid SS2 is not enough in itself to movethe 3-4 shift valve, so it is held in the position shown in FIG. 9B.

During light throttle operation, the 2-3 backout valve feeds the directclutch through the orifice 360. During heavy throttle operation, the 2-3backout valve is shifted to the right, thereby forcing supply oil to bedirected through the orifice 358.

FIGS. 10A and 10B show the valve system of FIG. 4 with the valves movedto the positions they assume for overdrive, fourth gear, part throttle,open-converter operation.

FIG. 10A shows the converter and bypass clutch control circuit. Asmentioned previously, the duty cycle of the modulated converter clutchsolenoid can be increased to force the bypass clutch control to itsdownward position when full clutch application (no slip) is desired.Modified pressure from the modulated converter clutch solenoid willeffect a controlled slip as explained previously. The duty cyclecommanded by the processor then is greater than zero, but it is lessthan 100%.

FIGS. 11A and 11B show the valve circuit of FIG. 4, but with the valveelements moved to the positions that correspond to drive range D, thirdgear, part throttle operation with a modulated converter clutch.

The 2-3 shift valve, the 3-4 shift valve and the overdrive servo is bestseen in FIG. 11B. Both the apply and release side of the overdrive servoare pressurized and the servo is released. As seen in FIG. 11B, the 3-4shift valve is downward. Solenoid pressure shifts the 2-3 shift valveand the 1-2 shift valve in an upward direction. As explained previously,this will cause the forward clutch and the direct clutch to be appliedas well as the intermediate clutch.

FIG. 12A and FIG. 12B show the valve elements of FIG. 4 in the positionsthey assume during first ratio operation in the manual low range with aclosed throttle and an open converter. The solenoid SS2 is released andthe solenoid SS1 is applied. Thus the 1-2 shift valve is moved downward.That forces the 2-3 shift valve also to move downward. Thelow-and-reverse servo is applied as pressure is distributed through the1-2 shift valve from the manual valve. The 2-3 capacity modulator valveis forced to its lower position and the orifice control valve is forcedto its upper position. The feed of fluid to the forward clutch isthrough orifice 380.

FIGS. 13A and 13B show the valve circuit of FIG. 4 with the valves movedto the positions corresponding to the second gear in the manual driverange with part throttle and an open converter. The overdrive servoapply chamber is pressurized at this time and the 1-2 shift valve ismoved upward, thus permitting pressure from the manual valve to bedistributed to the overdrive servo regulator valve and to the 1-2accumulator and through the orifice 306 to the intermediate clutch. Bothshift solenoids are turned off by the processor. The modulated converterclutch solenoid is turned off and the converter clutch is released. Thisis true also for first gear.

FIGS. 14A and 14B illustrate the circuit of FIG. 4, but the valves havebeen moved to the positions that they assume when the vehicle first isstarted and the manual lever is moved to the overdrive range positionand the vehicle is at rest. The valves in this condition control theengagement of the forward clutch. Soft engagement is assured by thefluid delivery from the orifice control valve through the flow controlorifice 380. The 2-3 accumulator also is pressurized to cushion theshift as fluid from the feed line 370 for the forward clutch isdistributed through the check valve 362 to the upper end of theaccumulator chamber.

FIGS. 15A and 15B illustrate the positions of the valves when they areconditioned for reverse drive and the reverse clutch is being engagedfollowing movement of the manual valve from the neutral position to thereverse position. The manual valve at this time delivers line pressureto the low and reverse servo and to the reverse clutch, the rate ofapplication of the reverse clutch being controlled by the orifice 472.Check valve 464 provides communication between the manual valve and thefeed line for the low and reverse servo. Shift solenoid SS1 is on andshift solenoid SS2 is off.

FIGS. 16A and 16B show the valve circuit with the valve elements inoverdrive range as a 1-2 shift is occurring with part throttle. Themanual valve then delivers fluid through the 1-2 shift valve, whichassumes the upward position under the influence of the 1-2 shift valvespring. The intermediate clutch is fed then by the 1-2 shift valvethrough the orifice 306. The pressure made available to the intermediateclutch acts on the lower side of the 1-2 accumulator piston 448. Withline pressure acting on the middle area of the piston 448, theapplication of the intermediate clutch is cushioned.

FIGS. 17A and 17B show the valve circuit with the valve elementsconditioned for the overdrive range as the valves are causing a 2-3shift with part throttle. Forward clutch pressure is distributed throughthe 3-4 shift valve, which assumes a downward position. It assumes adownward position also during a 1-2 shift. The 2-3 shift valve movesupwardly because solenoid pressure acts on the lower end of the 2-3shift valve. The shift solenoid SS2 is on at this time and the shiftsolenoid SS1 is off. Upward movement of the 2-3 shift valve holds the1-2 shift valve up because, as explained earlier, they are connectedmechanically, one with respect to the other. Overdrive release pressureis distributed through passage 328 from the 2-3 shift valve as explainedpreviously. The intermediate clutch is pressurized as feed line 302 isbrought into communication through the 1-2 shift valve with linepressure passage 206, which extends to the manual valve.

FIGS. 18A and 18B show the valves in the positions they assume inoverdrive range while a 3-4 upshift is taking place at part throttle.Line pressure is distributed through the 1-2 shift valve which, togetherwith the 2-3 shift valve, moves to the upward position because of thesolenoid actuating pressure. The overdrive servo regulator valve movesdownward thus establishing communication between the 1-2 shift valvepassage 294 and the apply side of the overdrive servo.

FIG. 18B shows the positions of the shift valve elements. Both shiftsolenoids SS1 and SS2 are applied. Both shift solenoid valve pressuresthen are sufficient when they act on the 3-4 shift valve to move the 3-4valve against the force of the valve spring for the 3-4 shift valve. The3-4 shift valve thus is able to direct fluid to the overdrive servoregulator valve through passage 440. This keeps the overdrive servoregulator valve from regulating, thus allowing full line pressure to bedistributed to the apply side of the overdrive servo. The 3-4 capacitymodulator valve 420 softens the 3-4 shift by regulating the pressure onthe release side of the overdrive brake band as fluid is displaced fromthe servo through the passage 328. The pressure on the forward clutch isexhausted through the check valve 414 and through the 3-4 shift valve atthis time.

FIGS. 19A and 19B show the valves in the positions they assume duringoverdrive operation as a 4-3 coasting downshift occurs with a closedthrottle. The backout valve 330 assumes a left-hand position asexplained previously because the electronic pressure control solenoidvalve output pressure is a minimum at that time. The 3-4 shift valvemoves downward because shift solenoid SS1 is deactivated and theremaining pressure at the lower end of the 3-4 shift valve is suppliedfrom the shift solenoid SS2 is insufficient to overcome the force of thevalve spring 400. This makes pressure available from the manual valve tothe forward clutch through the orifice control valve and the 2-3 backoutvalve through orifice 388.

FIGS. 20A and 20B show the valve elements in the positions they assumewhen the manual lever is in the overdrive position and a 4-3 downshiftis occurring with part throttle. The overdrive servo regulator valve 314cushions a 4-3 downshift with part throttle. The overdrive servoregulator valve is a simple regulator valve that modifies the pressuredelivered to the apply side of the overdrive servo. The magnitude of themodification depends upon the spring force for the regulator valve spool438.

FIGS. 21A and 21B show a valve diagram with the valves in the positionsthey assume when the transmission is in the overdrive range and a 2-1downshift is occurring with closed throttle. At this time, shiftsolenoid SS2 is off and shift solenoid SS1 is on. Thus, the 1-2 shiftvalve is forced downward. That keeps the 2-3 shift valve downward also.Since both sides of the overdrive servo are pressurized, the servo isreleased. The backout valve is shifted in a left-hand direction, andforward clutch pressure is distributed through the backout valve to theforward clutch through orifice 380.

FIGS. 22A and 22B show a valve diagram with the valve elements in thepositions they assume when the manual valve is in the overdrive rangeposition and a 3-2 downshift is occurring with closed throttle. Manualvalve pressure is distributed through the 3-4 shift valve, which assumesits downward position. Pressure then is distributed through the orificecontrol valve and to the forward clutch through the orifice 380. Thedirect clutch is exhausted, as is the lower chamber of the 2-3accumulator, through passage 316, through the 2-3 shift valve andthrough exhaust orifice 318.

The EPC Solenoid Valve

FIG. 25 shows electronic pressure control solenoid valve 114. Itincludes a variable force solenoid actuator (VFS) and a movable valvespool 480 with two spaced valve lands 482 and 484. The valve spool issituated in valve chamber 485 and is biased in an upward direction byvalve spring 486. Line pressure port 487, which receives regulated linepressure in passage 190 from the main regulator valve, communicates withsolenoid valve output pressure line 130 through the space between thelands 482 and 484 and through output port 489. Land 484 controls thedegree of communication between passage ports 487 and 489. Pressure inport 487 is transmitted to the upper side of land 482 through orifice491 and through central passage 493.

Solenoid windings 488 of the variable force solenoid actuator surroundsolenoid armature 490. A valve spring 492 is located in a centralopening of the armature and biases the armature in a downward direction.It is seated on the spring seat 494 secured to the solenoid housing. Thearmature is surrounded by flux washer 495.

When the solenoid windings are energized, the armature is moved upwardagainst the force of pilot spring 492. This decreases the seating forceof the poppet valve (on the bottom of the armature) against the poppetseat port 497 adjacent the bottom of the armature. This then decreasesthe restriction of port 497 and results in a shifting movement of thevalve spool 480, thereby restricting communication between passages 190and 130 and increasing the degree of communication between passage port489 and an exhaust port shown at 498. Port 498 is in directcommunication with the upper side of port 497. Land 484 registers withport 487 communicating with passage 190 and land 482 registers with theexhaust port 498.

A minimum current in the solenoid, which is commanded by the processor,will result in a maximum pressure output signal in passage 130. Amaximum current in the solenoid will result in a minimum signal pressurein passage 130. This results in a reduction in line pressure.

Modulated Converter Clutch Solenoid Valve

FIG. 26 shows the modulated converter clutch control solenoid valve 118.It receives a regulated input pressure from the solenoid regulator valvethrough passage 240. It distributes a variable output pressure to theconverter bypass clutch control valve-through passage 228.

The solenoid valve 118 includes a housing 498 which receives ball valveelement 500 located in valve chamber 502 formed in the housing 498.

Unlike the electronic pressure control solenoid valve of FIG. 25, whichis a variable force solenoid with an integral valve spool, the modulatedconverter clutch control solenoid of FIG. 26 is a pulse-width modulated(PWM) solenoid having an armature 504 that engages ball valve 500. Ballvalve 500 seats on a valve seat 506. The armature 504, when it isactuated in a downward direction, unseats the ball valve element thusproviding controlled communication between passages 228 and 240. Passage228 communicates with the upper side of the ball valve control orifice.

The solenoid is capable of operating at zero duty cycle when the ballvalve is closed, which causes the bypass clutch to be off. When it isoperated with 100% duty cycle, the ball valve is fully open and thepressure in passage 228 equals the pressure in passage 240. Controlledslip of the clutch is achieved when the duty cycle is between 0 and100%.

Solenoid windings 508 surround the armature 504. When the windings areenergized, the armature 504 moves downwardly thereby causing armatureextension 510 to unseat the valve 500. The extension 500 is movablypositioned in valve guide sleeve 512. Electrical connector 514 providesthe connection between the voltage source and the solenoid windings.

FIG. 23 is a chart that summarizes the engagement and the releasepatterns for effecting ratio changes as the shift solenoids SS1 and SS2are turned on and off. A separate chart section is provided for eachmanual valve position; i.e., OD, D, manual low and reverse. It should benoted that it is necessary only to change the state of a single solenoidin order to effect a ratio change between adjacent ratios duringoperation in any of the drive ranges selected by the manual valve. Forexample, when the manual valve is in the overdrive position, a ratiochange from the first ratio to the second ratio is achieved merely bychanging the state of shift solenoid SS1 from on to off. The state ofsolenoid SS2 remains unchanged. Similarly, a ratio change from thesecond ratio to the third ratio or from the third ratio to the secondratio is effected merely by changing the state of solenoid SS2 while thestate of solenoid SS1 remains unchanged. Ratio changes between the thirdratio and the fourth ratio are obtained merely by changing the state ofsolenoid SS1 while the state of solenoid SS2 remains unchanged. Thischaracteristic makes it unnecessary to synchronize the actuation ordeactivation of one solenoid with respect to the activation ordeactivation of the other solenoid. This simplifies the calibration ofthe transmission and improves the reliability of the control system.

FIG. 24 shows the shift schedules for a preferred embodiment of thetransmission. The shift schedules, which are stored in ROM, areillustrated in the chart of FIG. 24 by plotting throttle opening for theengine versus vehicle speed for each ratio change. A separate chart isshown for upshifts and downshifts between the first and second ratios,for the upshifts and downshifts between the second and third ratios andthe upshifts and downshifts between the third and fourth ratios.

The information of FIG. 24 is stored in the ROM portion of themicroprocessor memory and is fetched from memory in response to thethrottle position and vehicle speed signals received by themicroprocessor. For any given vehicle speed and throttle opening, ashift will be indicated in accordance with the schedule of FIG. 24. Theprocessor then will develop an output signal that is distributed to theshift solenoids SS1 and SS2 to effect the appropriate ratio change tosatisfy the correct vehicle speed and throttle opening.

It should be noted that during fourth ratio operation, both solenoidsSS1 and SS2 are on. Thus, leakage is avoided through the shift solenoidsduring the majority of the operating time when the transmission is inthe fourth ratio. Further, as previously discussed, the ratio changesbetween adjacent ratios requires only activation or deactivation of asingle shift solenoid, thus eliminating the need for providingsynchronous operation of the shift solenoids. The response time thatresults from this is very short and the shift valve movement is precise.As soon as the shift command is issued, an immediate response of theshift valve to the change of state of the shift solenoid will occur. Forexample, on a 4-3 downshift, when the shift solenoid SS1 goes off, fluidflow immediately passes from the shift solenoid pressure passage 408through the shift solenoid SS1. The orifice 242 associated with theshift solenoid SS1 has limited flow capability. That limited flowcapability will not delay a shift of the 3-4 shift valve in a downwarddirection. The same is true for orifice 244 on a 3-2 downshift when theshift solenoid SS2 is turned off. The exhaust of fluid from passage 246will not be delayed because of the limited flow capability of orifice244.

The response time is most important in a 4-3 downshift and a 3-2downshift and the circuitry is fastest in those shifts. The solenoidshave ample capacity to displace fluid to exhaust.

On every downshift, except the 2-1 downshift, the strategy commands theassociated shift solenoid to be turned off. This results in a rapidfall-off of the solenoid pressure and a quick response of the shiftvalve movement.

The 3-4 shift valve is hydraulically connected directly to the manualvalve. Thus, when the manual valve is moved from the OD position, linepressure is sent through passage 108 directly to the top of the 3-4shift valve spool 412. This forces the 3-4 shift valve to assume itsdownshift position regardless of the solenoid state and regardless ofwhether pressure exists in passages 408 or 406. Since this is the case,the forward clutch will be being pressurized in all instances when themanual valve is pulled out of the OD position to another forward motionposition because the resulting downshift movement of the 3-4 shift valveelement 412 will directly connect line pressure passage 410 to theforward clutch through passage 370. The forward clutch is an importantclutch to be maintained in an applied condition in the event of afailure since it is needed for first, second and third ratios.

When the 3-4 shift valve is forced to its downshift position as themanual valve is pulled from the OD position, the pressure made availableto the 3-4 shift valve through passage 108 is distributed also to thetop of the overdrive servo regulator valve through passage 440 asexplained. This will insure that the overdrive servo apply pressure willbe at a maximum valve. Thus, the regulating action of the overdriveservo regulator is overruled under these conditions.

Control Circuit Components

Shown in FIG. 22C is an exploded view of the cast valve body andassociated components. The main valve body, shown at 516, comprisesmultiple cavities and valve bores for accommodating the valve elementsdescribed with reference to FIG. 4. It is a die casting with channelsthat conduct oil from one location in the valve system to the other. Thechannels are cast within the valve body during a die casting operation.

The valve body is covered with a cover plate 518 on one side thereof andby another cover plate 522 on the opposite side thereof. A gasket 532 issituated between the cover plate 522 and the surface of the maintransmission casting to which the valve body is secured. A second gasket530 is between cover plate 518 and the valve body 516. A transmissionoil pan 524 covers the assembly and is bolted to the main transmissionhousing.

Dowel bolts 526 and 528 align and cover plate 518 and its gasket 530 tothe valve body 516 and, during assembly, align gasket 520 and plate 522to valve body 516. The dowel bolts also align the main control assemblyto the housing casting.

In FIG. 22D there is shown a plan view of the subassembly after thecover plate 518 is bolted to the main valve body 516 by the dowel bolts526 and 528.

FIG. 22F is a side view of the subassembly of FIG. 22D. It shows themain valve body 516 and cover plates 518 and 522. Gasket 530 is seenlocated between the cover plate 518 and the valve body 516. Similarly,gasket 520 is seen located between the cover plate 522 and the valvebody 516.

When the dowel bolts are assembled, a dowel pin extension of the boltsprotrudes through dowel openings in the cover plate 522 and in thegasket 520. In the case of bolt 528, which is shown in FIG. 22E, thedowel pin extension is indicated at 534. These dowel pin extensions areadapted to be received in dowel openings formed in the lower portion ofthe transmission housing against which the cast valve body 516 isassembled.

The dowel bolt 528 is illustrated in detail in FIG. 22E. It includes athreaded shank 536, which is threadably received in a threaded openingformed in the cast valve body. The bolt head 538 clamps the cover plate518 and the gasket 530 on one side of the valve body. The dowel pinextension 534 is adapted to be received in dowel openings formed in thegasket 520 and the cover plate 522 located on the opposite side of thevalve body.

The dowel opening in cover plate 518 shown in FIG. 22D through which theassociated dowel pin extension extends is elongated in the direction ofthe major axis of the valve body. The opening in the direction of themajor axis is greater than the diameter of the dowel pin extension andthe width of the opening in the direction of the minor axis of the valvebody, except for tolerance, is equal to the diameter of the dowel pinextension. On the other hand, the opening for the dowel pin extensionfor dowel bolt 528 is round and, except for tolerance, is equal indiameter to the diameter of the dowel pin extension.

The elongated opening for the dowel bolt 526 accommodates axial stack-uptolerances while permitting precise registry of the valve plate 518 at aprecise location with respect to the valve body 516. Further, the dowelpin extensions 534 of the dowel bolts permit precise registry of thecover plate 522 and the gasket 520 to the valve body with respect to thepressure passages and also to the fluid conduits formed in the portionof the main transmission housing to which the valve body is assembled.

The shank portion 540 of the dowel pin 528 is provided with a diameterthat, except for tolerances, matches the diameter of the cover plate ofthe valve assembly through which it extends.

Bypass Clutch Control

The modulated converter bypass clutch solenoid establishes a controlledslip in the converter as explained previously. To assist in thedescription of the improved bypass clutch strategy of the presentinvention, a known bypass clutch described in U.S. Pat. No. 5,029,087now will be made. For a complete description of this known controlstrategy, reference may be made to U.S. Pat. No. 5,029,087. For purposesof the present description, however, reference will be made only toFIGS. 28 and 29 to explain, in general terms, the control strategy forachieving the necessary slip to accommodate the particular operatingconditions that are detected by the engine and driveline sensors. Thisinvolves a determination of the desired slip indicated in the chart ofFIG. 29 and comparing it to the actual slip that is measured by thesensors.

As explained previously, the actual slip is equal to the engine speedminus the turbine speed. The turbine speed, however, must be computedsince there is no turbine speed sensor for the presently describedembodiment of the invention. This is done by the processor by measuringoutput shaft speed and by correcting the output shaft speed by the gearratio that happens to be present at any given instant. The differencebetween the desired slip and the actual slip is indicated as an error Efor each of the times indicated in FIG. 29; for example, time T0, T1 andT2. This error is calculated during each background loop of themicroprocessor.

As seen from FIG. 29, the error progressively decreases as the targetvalve for slip is approached. This target value is a value stored in theROM portion of the microprocessor memory. Its value is fetched frommemory in accordance with the normalized turbine speed that is computedas explained above, and with the normalized throttle position.

The control strategy for determining desired slip is illustrated in theflow chart of FIG. 28. During the process of determining desired slip,the processor will sequence through the various stages of FIG. 28following the setting of the desired slip equal to the absolute valuefor the converter slip. This is done as indicated at action block 542 ofFIG. 28. The processor then determines which gear ratio is in place. Ifthe gear ratio is greater than third gear, a command is issued to lookat the table in memory where target slip is stored. This is done at step545. The inquiry with respect to whether the gear ratio is greater thanthree is made at step 546. That target value, after having been placedin the target slip register, is used for purposes of determining desiredslip. The desired slip is obtained using the equation:

Desired slip=Actual slip-(PCDEC "x") times (Actual slip-Target slip),where PCDEC "x" is the value that appears at 548, 550 or 552 in FIG. 28.

Referring again to FIG. 29, a plot is made of slip versus time. Curve554 is a plot showing the actual slip versus time and plot 556 shows thedesired slip versus time. Plot 556 is a result of the calculationdescribed above using the three error values described previously; i.e.the error (E₋₋ TO) currently measured by the slip controller, the error(E₋₋ T1) determined in the previous background loop and the error (E₋₋T2) determined in the second previous background loop. These errorvalues are indicated in FIG. 29.

In FIG. 29, the target slip is indicated by the symbol T. The value ofthat target slip, as mentioned previously, is determined by the throttleangle and turbine speed that exist for any particular gear ratio. Asmentioned previously, that information is obtained from a table inmemory.

As seen in FIG. 29, the desired slip is calculated for each backgroundloop so that the value of the absolute slip approaches asymptoticallythe value of the target slip T.

The procedure used to calculate desired slip is graphically illustratedin the flow chart in FIG. 28. The routine begins as mentioned previouslyat stage 542 where the value of the absolute slip is determined basedupon the actual slip value. Having established an absolute slip value,the routine proceeds to stage 546 where as mentioned previously aninquiry is made with respect to the gear ratio that is determined by thegear ratio sensor. If the gear ratio is greater than the third ratio,the sequence proceeds to stage 544. If the ratio is the fourth ratio atthat time, the appropriate stored value of the target slip is read frommemory. On the other hand, if the gear ratio is that gear ratiocorresponding to third gear, another target value is read at stage 545.That occurs if the inquiry at stage 546 is negative. In either case, theroutine will proceed to stage 547, where an inquiry is made with respectto the gear ratio in place.

If the gear ratio is that gear ratio corresponding to fourth gear, theslip reduction factor PCDEC4 is stored in the temporary storage registeras shown at stage 548. If the inquiry at stage 547 is negative, theroutine proceeds to stage 549 where another inquiry is made to determinewhether the gear ratio is that gear ratio corresponding to third gear.If the answer to that inquiry is positive, slip reduction factor PCDEC3corresponding to that gear ratio is fetched from memory and stored inthe temporary storage register as shown at stage 550. If the inquiry atstage 549 is negative, the routine proceeds to stage 551 where adetermination is made as to whether the gear ratio is that gear ratiocorresponding to second gear. If the answer to that inquiry is positive,a temporary holding register receives still another slip reductionfactor PCDEC2 from memory. That stage is indicated at 552. If theinquiry at 551 is negative, the routine proceeds to stage 553 whichcauses the temporary holding register to receive a new data item(PCDEC1) from memory.

The routine then passes to the final stage 554 where a desired slipcalculation takes place using the temporary register data in the formulapreviously described. In other words, desired slip is equal to theabsolute slip minus a portion of the difference between the absoluteslip and the target slip for the particular background loop in place. Atthe end of each background loop, the previous error E₂ is set equal toE₁ and the previous error E₁ is set equal to the current error E₀. Thisupdates the information for each background loop so that a new error canbe calculated for the next loop.

After the error is determined, the duty cycle is determined using thedesired converter slip calculated in stage 554.

If the clutch is in the unlock mode, the duty cycle is zero. If theclutch is in the stroke mode, the duty cycle is adjusted as a functionof the throttle position to provide insufficient hydraulic pressure tocause the clutch to continue to be disengaged, but such that anyadditional pressure will cause the clutch to engage. Thus, the clutch ismaintained at a so-called incipient engaged (stroked) condition.

During the engage mode, however, the duty cycle calculation, using aclosed loop technique, provides an incremental gain PID controller withinformation to adjust the pulse width modulated duty cycle to obtain thedesired slip. This procedure starts with the absolute PID formulas:

    Duty cycle output new=K.sub.p (E.sub.0)-K.sub.d (E.sub.0 -E.sub.1)/T.sub.o +K.sub.i (E.sub.0 *T.sub.o +E.sub.1 *T.sub.1. . . +E.sub.n *T.sub.n)

    Duty cycle output old=K.sub.p (E.sub.1)+K.sub.d (E.sub.1 -E.sub.2)/T.sub.1 +K.sub.i (E.sub.1 *T.sub.1 +E.sub.2 *T.sub.2. . . +E.sub.n *T.sub.n)

In the equations above E₀ is the current error, E₁ is the previous errorand E₂ is the second previous error. T₀ is the time of the currentcontrol loop between the current and the previous error readings, T1 isthe time of the previous control loop between the previous and thesecond previous error readings. T₂ is the time between two successiveearlier error readings. The constant K_(p) is a proportional gainconstant, the constant K_(d) is a derivative gain constant and theconstant K_(i) a integral gain constant. The output change equals theoutput-new formula minus the output-old formula. That value is computedin accordance with the following formula:

    Output-change formula=K.sub.p (E.sub.0 -E.sub.l)+K.sub.d [(E.sub.0 -E.sub.1)/T.sub.0 +(E.sub.2 -E.sub.l)/T.sub.1 ]-K.sub.i *E.sub.0 T.sub.0

As mentioned previously, there are advantages in causing the desiredslip for the bypass clutch to approach its target value quicklyconsistent with the ability of the bypass clutch to reduce or eliminatenoise, vibration and harshness in the driveline, particularly inertiatorque disturbances due to transient torque fluctuations. For thispurpose, the desired slip determination for the present invention isdifferent than the desired slip calculation described above withreference to the control strategy of U.S. Pat. No. 5,029,087. This willbe explained first with reference to FIG. 26B, where slip is plottedagainst time in a manner similar to the plot of FIG. 29. It should benoted from FIG. 26B, however, that the desired slip, which isrepresented by the line shown at 556, is set equal to the absolute slipor actual slip at point 558. The absolute slip or actual slip isrepresented by the curve shown at 560. This corresponds to the curve 554of FIG. 29.

The absolute slip, as mentioned earlier, is determined by the processorand is equal to the difference between the sensed engine speed and theturbine speed. The turbine speed, as mentioned earlier, is a speeddetermined by the processor by measuring output shaft speed andmultiplying that output shaft speed value by the gear ratio that is inplace at the instant the speed is measured. The transmission of FIGS. 1and 2 does not have a turbine speed sensor, so this speed calculationmust take place to achieve an accurate indication of converter slip.

The difference between the absolute slip and the desired slip duringeach background pass is indicated as an error. The error during abackground pass or loop n is identified as error E₋₋ T₂, whichcorresponds to error E₋₋ T₂ of FIG. 29. The error for the nextbackground pass or loop n+1 is identified as error E₋₋ T₁, as in thecase of FIG. 29. The error for the current background pass or loop n+2is identified as E₋₋ T₀, as in the case of FIG. 29. A duty cycle iscalculated in the manner previously described corresponding to each ofthese computed errors during each control loop. This results in thedecay of the absolute slip toward a target value T, seen in FIG. 29.

The desired slip is a computed value resulting in a ramped relationshipas shown at 556. It intersects the target value T at point 562, seen inFIG. 26B. The ramped relationship shown at 556 is determined by a sliprate modifier or constant multiplied by the initial desired slip valueindicated at 558. The magnitude of that modifier depends upon the gearratio that is in place. For an understanding of this relationshipsreference now will be made to FIG. 26A which shows the control strategyflow chart for the desired slip determination. This is a PID closed loopclutch control.

In FIG. 26A, the beginning of the slip desired determination strategybegins with an inquiry made at step 564. If the engage flag in theclutch control routine is on, the beginning of a clutch engagement isindicated. If the inquiry indicates that the clutch has begun to engage,the slip desired is set equal to the absolute slip as indicated ataction block 566. After the slip desired is set equal to the absoluteslip, as indicated at 558 in FIG. 26B, the routine will proceed toinquire at step 568 whether the gear ratio that has been commanded isfourth ratio. If the fourth gear ratio has been commanded, the routinewill proceed to action block 570, where a slip rate factor for rampingthe desired slip relationship is fetched from memory. This factor iscalled SR4 in FIG. 26A. The routine then will proceed to calculate adesired slip. The routine then proceeds to step 572 as earlierindicated.

If the answer to the inquiry at step 574 is negative, the slip rate thenis determined at action block 578 where a slip rate value SR2 isobtained. That value is consistent with the ramp that is appropriate forsecond gear ratio. There is no slip rate called out for the first ratiosince the routine of FIG. 26B is not carried out for the first ratio.The converter acts only as an open converter during first ratiooperation.

If the routine proceeds to action block 572, the desired slip iscalculated by subtracting the desired slip value determined at actionblock 566 and subtracting the slip rate value located at either actionblock 570, action block 576, or action block 578, depending upon thegear ratio that is in place.

The routine then proceeds to action block 580 where an inquiry is madeto determine whether the desired slip value has become equal to thetarget value T. That would indicate that the intersection point 562 ofFIG. 26B has been reached. If the inquiry at 580 results in anaffirmative answer, the desired slip at action block 582 is determinedto be equal to target slip and no further decay in the desired slip isallowed. The routine then proceeds back to the beginning and the routineis repeated in the next background loop.

When the driveline variables have stabilized, a point will be reached attime value 584, shown in FIG. 26B, where a command is issued to enter aso-called "hardlock" operating mode. The need for maintaining a targetslip is not deemed to be necessary after driveline stability isachieved.

When the processor detects that point 584 has been reached, the desiredslip then will be ramped down as indicated at 586 in FIG. 26B. A rampvalue appropriate for the hardlock condition is fetched from memory atthat point and is used to establish the ramp slope for the line 586. Themagnitude of the target slip T at the point 584 is multiplied by thehardlock modifier. The absolute slip after the hardlock mode has beeninitiated in this fashion is different than the desired slip line 586.As in the case described earlier with respect to absolute slip curve 560and desired slip curve 556, the processor will determine during eachbackground loop the difference between the absolute slip and the desiredslip. The duty cycle then is determined for each background loop in themanner previously described. When the absolute slip reaches the targetvalue at point 588, the clutch will be fully locked. This signals theprocessor to increase the magnitude of the duty cycle to 100%, thuscausing the clutch to be fully engaged.

Shown at 590 in FIG. 26B is a duty cycle plot over time during theengagement routine for the bypass clutch. The duty cycle value for point558 on the desired slip and absolute slip curves is indicated at 592.The magnitude of the duty cycle will gradually be increased for eacherror reading. This is done in the manner previously described. Thus,the duty cycle will increase as shown at 594 in FIG. 26B and will becomerelatively constant as shown at 596 after the absolute slip reaches thetarget value T.

When the hardlock mode is entered at point 584 in FIG. 26B, theprocessor will command a higher duty cycle as shown at 598. As in thecase of the computation of duty cycle corresponding to absolute slipcurve 560 and the desired slip relationship shown at 556, a duty cyclecalculation is made for each error detected during each background pass,while the desired slip is ramping as shown at 586. This is done in themanner previously described with respect to ramp portion 556. The resultof this duty cycle calculation is a decrease in the absolute slip untila zero slip value is reached at 588.

When the point 588 is reached, the processor commands an increase in theduty cycle as shown at 600 in FIG. 26B until a 100% duty cycle isreached as shown at 602.

FIG. 26E shows the electronic control strategy for the hardlock mode inflow diagram form. At the beginning of the hardlock mode, upon enteringstage 584, the processor, as it continuously monitors speed ratio, willdetermine whether a target slip of zero has been reached. This is doneat step 604. If the processor detects that the slip target has not beenreached, the routine will proceed directly to action block 606 where theclosed-loop PID control is continued. In the next background loop, thetest for target slip again is made at step 604. If the inquiry at step604 finally is positive, an inquiry then is made at step 608 todetermine whether a shift is taking place. If a shift indeed is takingplace, the routine will proceed to action block 610 where entry into thehardlock mode is inhibited. If a shift is not taking place, the routinewill proceed to action block 612 where the slip is controlled, asdescribed previously with reference to the absolute slip curve 614 inFIG. 26B and with reference to desired curve ramp 586. That wouldindicate that the point 588 has not yet been reached.

An inquiry then is made at step 616 to determine whether the absoluteslip has finally stabilized at the zero slip line. If it has not, theroutine then will return as shown in FIG. 26E to action block 612, wherethe closed loop PID control continues.

If the slip finally has stabilized at zero, the routine proceeds toaction block 618 and fetches from memory a multiplier for the duty cycleso that the duty cycle is increased to 100% as indicated at 602.

The behavior of the clutch during shifts can be described best withreference to FIG. 26C. As mentioned earlier, the transmission describedhere does not include a turbine speed sensor. Thus, it is necessary todetermine the turbine speed by taking into consideration the gear ratiothat is in place at the instant a slip measurement is made. In the plotof FIG. 26C, engine speed is plotted as shown at 620. During the periodof acceleration of the vehicle, the engine speed rises. If it is assumedthat the vehicle is in third gear, the vehicle speed will increasegradually as indicated by the rising portion 622 of the speed curve 620.The gear command value is designated at 624. At time 626, a command ismade to shift the transmission from third ratio to fourth ratio.

As indicated by shift command line 628, the speed ratio across theconverter continuously is monitored. The speed ratio value at a timeprior to the commanding of the shift at 626 is indicated at 630. Thespeed ratio in the particular case of a transmission that does not havea turbine speed sensor is determined by multiplying the output shaftspeed by the gear ratio in place, divided by engine speed. When a shiftfrom the third ratio to the fourth ratio is commanded at time 626, thespeed ratio that is monitored drops from a value shown at 630 to a valueshown at 632 in FIG. 26C. At that time, the duty cycle is frozen orconditioned for an unchanging output as indicated at 634. The duty cyclefor the modulated converter clutch solenoid valve is plotted as shown at636. At a time prior to the commanding of the shift at time 626, theduty cycle fluctuates as shown by the wavy line 638. The duty cycleremains constant, however, in the freeze zone 634.

The shift actually will begin at a time 640 following the command at626. The beginning of the shift is indicated when a delta speed ratio642 is detected by the speed sensors. Actually, the beginning of theShift will occur slightly in advance of point 644 in FIG. 26C. Theachievement of the delta speed ratio will confirm that a shift actuallyhas begun. After that confirmation is made, the open-loop controltechnique described earlier will begin at time 646.

The beginning of the open-loop control coincides with a flag thatinstructs the pointer for the microprocessor to fetch from memory amultiplier for the duty cycle. That multiplier is a value less thanunity. When the duty cycle value 634 is multiplied by the multiplier,the effective duty cycle decreases as shown at 638.

The speed ratio, which as mentioned earlier, is continuously monitored,will detect an increased delta speed ratio. When that delta speed ratiois large enough to indicate that the shift is ended, as shown at point648 in FIG. 26C, the duty cycle is increased to a value substantiallythe same as the value that existed at 634. This is indicated at 650.

At time 652, following the completion of the shift, the controller willreturn to the closed-loop PID control and initialize again the PIDclosed-loop control strategy discussed earlier. Following the completionof the shift, the engine speed again rises in normal fashion as vehicleacceleration continues. This is shown at 654. Between the time at 640when the shift begins and the time at 652 when the shift ends, theengine speed decreases at indicated at 656.

Shown in FIG. 26D is a chart of data showing data stored in memory whichrefers to the open loop function that begins at 640 in FIG. 26C which isthe time the shift ends at 652 in FIG. 26C. The delta speed ratio, whichis obtained by continuously monitoring the engine speed and the outputshaft speed, is plotted on the horizontal axis and duty cyclemultipliers are plotted on the vertical axis in FIG. 26D. At thebeginning of the open loop control, the multiplier, in the example shownin FIG. 26D, is about 0.8. This is shown by the horizontal line 658.Later in the open loop control, as determined by expiration of acalibrated timer, the duty cycle is increased as shown in FIG. 26C, ashigher multipliers are pulled from the ROM portion of the memory. Thisis shown at 660 in FIG. 26D. Finally, when the delta speed ratio reachesthe point corresponding to point 648 in FIG. 26C, the multiplier is at amaximum value, as shown at 662 in FIG. 26D.

Shown in FIGS. 30 and 30A is a flow chart showing the shift modulationlogic that occurs during a shift interval. This further explains thebehavior of the clutch during the shift, which is graphicallyillustrated in FIG. 26C.

The logic of FIGS. 30 and 30A comprises a series of logic steps that arecarried out by the processor during a background control loop. At thebeginning of the control loop, the processor will determine at step 670whether a shift has been commanded. In the case of the shiftdiagrammatically represented by the plot of FIG. 26C, the shift begins,as explained previously, at time 626. The shift in the case of FIG. 26Cis a 3-4 upshift. After the command of the shift, the shift will occurlater at time 640.

If it is determined at step 670 that a shift has been commanded, theprocessor will set a shift command flag and then proceed to step 672.The processor at step 672 will initiate four calculations, separatelyidentified in FIG. 30 by numerals 1, 2, 3 and 4. In calculation 1 ofstep 672, the processor will calculate the speed ratio for the new gear.This is done by multiplying the filtered engine speed by the gear ratiothat has been commanded divided by the engine speed at the instant theshift is commanded. In calculation 2 of step 672, the processor willrecord the duty cycle that exists at the time of the shift command. Thisis the duty cycle indicated at point 674 of FIG. 26C. Calculation 3 ofstep 672 calls for the processor to record, in temporary storage memory(RAM), the speed ratio at the instant the shift is commanded. This isthe speed ratio indicated at point 676 of FIG. 26C. In calculation 4 ofstep 672, the processor will set a flag that indicates that the upshiftmodulation control is in progress.

The next step in the routine, as indicated at step 678, involves aninquiry as to whether the upshift that has been commanded is an upshiftto the fourth ratio as in the case of FIG. 26C. If in fact the upshiftcommanded is a shift to the fourth ratio, the routine then will proceedto action block 680 of FIG. 30 where a calculation is made to determinethe speed ratio that must exist to provide an indication that the shiftis complete. This value, after being calculated, is stored in RAM forfuture reference. If in fact the inquiry at step 678 is negative, theroutine will proceed to action block 682 where a similar calculation ismade to determine the speed ratio that would indicated that the upshiftis complete.

If the upshift is not a shift to fourth, then the upshift, as seen inaction block 682, is an upshift to third. In the calculations at step680 or at step 682, use is made of calibration constants that arefetched from memory depending upon the gear ratio that is involved inthe upshift. A different constant is established for each gear ratio. Ineither case, the speed ratio that would indicate that the shift hasbegun equals the speed ratio that exists at the beginning of the shift;namely, the speed ratio at point 676 in FIG. 26C in the case of a 3-4upshift, times the calibration constant appropriate for that upshift.

Following the calculation at step 682, the routine will proceed to step684 where a calculation is made of the speed ratio that would indicatethat the shift has begun. This is done by assuming that the speed ratiois at the value at point 676. That value is determined by the softwareto be the new speed ratio, even though the shift has not yet occurred.Simultaneously, the duty cycle has been frozen, as indicated at 634 inFIG. 26C, beginning at point 674. Since the duty cycle is frozen andsince the speed ratio is assumed to be that ratio indicated at point 676in FIG. 26C, the only variable that can effect a change in speed ratiois a change in engine speed. This is indicated in FIG. 26C where theengine speed changes from the value at 688 to the value at 690. When thevalue at 690 is reached, a delta speed ratio, indicated in FIG. 26C,will equal a delta speed ratio stored in memory, which is a calibrationconstant.

As a result of the calculation of the speed ratio at point 686, theprocessor will reset the flag, which will indicate that the shift hasbegun. The routine then will proceed to make an inquiry at step 692 ofFIG. 30A as to whether the shift actually has begun. If it has notactually begun, the routine will return, as shown at 694, to thebeginning and the inquiry is repeated. This occurs during a onemillisecond repeater or interrupt loop. This is a foreground controlloop that is carried out during the longer background loop, which may beabout 40 milliseconds. The inquiry at 692 will be made repetitivelyuntil a confirmation is reached that the shift has begun. The routinethen will proceed to action block 694, which sets the flag to indicatethat the shift has begun.

The next step in the routine occurs at 696 of FIG. 30A where aconfirmation is made as to whether the upshift to fourth gear has beencommanded. If it has, the duty cycle is calculated at action block 698by calling from memory a multiplier constant from a table stored inmemory corresponding to FIG. 26D where the delta speed ratio is plottedagainst a multiplier value. Initially, the multiplier may be about 0.8as indicated in FIG. 26D. This results in a reduction in the duty cycleas shown at 636 in FIG. 26C. The magnitude of the duty cycle in the openloop control portion of the shift is determined by the multiplier valuesince the duty cycle is calculated by multiplying the duty cycle at thestart of the shift by the multiplier value. As indicated in FIG. 26D,the multiplier will increase as shown at 660. This results in anincrease in the duty cycle as shown in FIG. 26C, until a duty cycleshown at 650 is reached.

If the inquiry at 696 is negative, that would indicate that a 2-3upshift has been commanded rather than a 3-4 upshift. In that case, theroutine will proceed to action block 700, rather than to action block698. There a calculation occurs for a 2-3 upshift to determine theappropriate value for a duty cycle for open loop control. The magnitudeof the multiplier for calculating the open loop duty cycle is differentin the case of a 2-3 upshift than for a 3-4 upshift.

The shift is complete when the delta speed ratio is large enough toindicate the end of the shift, as shown at point 648 of FIG. 26C.

As indicated earlier, an inquiry is made at step 670 as to whether theshift is being commanded. If it is commanded, the four calculationsindicated in action block 672 are carried out. This provides a"snapshot" of the variables that exist at the instant a shift iscommanded. If the inquiry at 670 is negative, an inquiry is made at step702 of FIG. 30 to determine whether the upshift modulation control is inprogress. If it is, the routine will proceed to action block 704, wherethe delta speed ratio is calculated. The routine then will return tostep 706 where an inquiry is made to determine whether the shift hasbegun by checking the state of the flag that was set at step 694. Ifthat flag has not been set, the routine then will proceed to the nextcontrol module at the end of the routine in FIG. 30A. If the shift hasin fact begun, the routine will make an inquiry at step 708 to determinewhether the shift that has been commanded is an upshift to the fourthratio. If it is not, the routine then will proceed to step 710 todetermine whether the shift has ended. That is, it is determined whetherthe delta speed ratio is greater than the value indicated at points 648of FIG. 26C. If it is greater than that value at point 648, the routinethen will proceed to the exit at 712, where the upshift modulationcontrol ends.

If the inquiry at step 708 is positive, again a test is made, as shownat 714, as to whether the shift has ended. This is done by determiningwhether the delta speed ratio is greater than the value indicated at648. If that value is reached, the routine ends as discussed above. Ifit has not been reached, the routine then will return to the step 696previously described where an inquiry again is made as to whether theupshift to the fourth gear is being commanded. If it is, the open loopcontrol duty cycle is calculated as described previously. If it is not,a corresponding determination of open loop duty cycle for a 2-3 upshiftis carried out at step 700 as previously described.

During tip-in or tip-out operation of the transmission, which occursduring a rapid advancement or a rapid relaxation of the engine throttle,a special tip-in, tip-out logic is used to decrement the bypass clutchduty cycle while the tip-in or tip-out is present. When the processordetects that a tip-in or a tip-out condition is in place, the PIDcontrol or the hardlock mode is temporarily suspended.

When the tip-in or tip-out condition is completed, a timer in theprocessor is used to control when the normal closed loop controlfunction is resumed. The time that is calibrated for the closed loopcontrol or the hardlock mode to return can be chosen as desired byappropriately calibrating the timer.

If the temperature sensor 110 of FIG. 3A indicates that the operatingtemperature is greater than a desired predetermined level, the processorwill command a duty cycle of 100%. This is a so-called "hot-lock"converter bypass clutch operating mode. This overrules the duty cyclecalculation, described above, that is appropriate for the modulatedconverter clutch slipping mode. The duty cycle calculation is overruledwhen the temperature reaches its maximum limit. The result of thatprocedure is a duty cycle output that causes the clutch to be fullyengaged whenever the maximum permissible temperature level is beingsensed by temperature sensor 110 seen in FIG. 3A.

Having described a preferred embodiment of the invention, what isclaimed and desired to be secured by U.S. Letters Patent is:
 1. In abypass clutch for a hydrokinetic torque converter for an automatictransmission and engine;a clutch member connected to said engine, aconverter having a friction surface adapted to cooperate frictionallywith said clutch member; a fluid pressure operated means for controllingthe clutching capacity of said bypass clutch whereby clutch slip may beeffected for a variety of engine and transmission conditions;electronically operated valve means for controlling the pressure appliedto said fluid pressure operated means, said valve means being connectedto and being responsive to an electronic processor adapted for controlfunctions in repetitive control loops and having control logic thatresponds during each control loop to changes in operating conditions ofsaid transmission and engine; means for determining a target slip foreach set of said operating conditions; means for determining the actualslip of said bypass clutch; means for determining a desired slip in aclutch slipping mode at the beginning of clutch engagement bymultiplying said actual slip during each control loop of said processorby a factor less than unity whereby said desired slip decreases until itequals said target slip, said actual slip thereby being made equal totarget slip, and means for establishing a hardlock mode after saidactual slip equals target slip whereby the torque flow path through saidconverter is bypassed.
 2. A bypass clutch as set forth in claim 1wherein said means for establishing a hardlock mode comprisesestablishing a new desired slip relationship at the instant hardlockmode begins by modifying actual slip at that instant to a calculatedamount less than actual slip, said actual slip thereby decaying untilactual slip is zero.
 3. The combination as set forth in claim 1 whereinsaid transmission has gearing with multiple gear ratios, said turbinebeing connected to driven portions of said gearing, means for detectingthe driven speed of said driven portion, means for continuouslymonitoring speed ratio of said converter as a function of gear ratio,said detected driven speed and said engine speed;the monitored speedratio decreasing upon a command of a shift to a higher speed gear ratio,means for detecting a delta speed ratio following said shift command toprovide an indication of the time a ratio change begins following saidshift command; means for modifying said clutching capacity to effect adecrease in clutch capacity and an increase in actual slip driving saidratio change; and means responsive to the attainment of an increase insaid delta speed ratio that is sufficient to indicate the end of saidratio change whereby said clutching capacity controlling means resumesits clutch control.
 4. In a friction bypass clutch for a hydrokinetictorque converter comprising an impeller, a turbine and an impeller shellenclosing said impeller and said turbine, an engine being connected tosaid impeller shell;a clutch plate within said impeller shell connectedto said turbine, said clutch plate defining with said shell a controlpressure chamber; control passage means communicating with said controlpressure chamber for distributing a clutch capacity determining controlpressure to said chamber, bypass clutch control valve meanscommunicating with said control pressure means, bypass clutch solenoidvalve means communicating with said clutch control valve means;electronic processor means for establishing a voltage signal for saidsolenoid valve means including means for sensing engine torque, meansfor sensing vehicle speed, means for computing turbine speed from saidvehicle speed, means for establishing a target slip based upon theengine torque and turbine speed values, means for measuring the actualslip of said clutch, means for determining desired slip by reducingactual slip by a predetermined amount until it equals said target slip,means for determining slip error equal to the difference between actualslip and desired slip, and means for establishing a control signal thatis established as a function of a first factor proportional to sliperror and to a second factor that is determined by a derivative of saiderror; means for subjecting said solenoid valve means to the controlsignal of said control signal establishing means; and means forestablishing a hardlock mode after said actual slip equals target slipwhereby the torque flow path through said converter is bypassed.
 5. Thecombination as set forth in claim 4 wherein said means for establishinga hardlock mode comprises establishing a new desired slip relationshipat the instant hardlock mode begins by modifying actual slip at thatinstant to a calculated amount less than actual slip, said actual slipthereby decaying until actual slip is zero.
 6. The combination as setforth in claim 4 wherein said means for establishing a hardlock modecomprises establishing a new desired slip relationship at the instanthardlock mode begins by modifying actual slip at that instant to acalculated amount less than actual slip, said actual slip therebydecaying until actual slip is zero.
 7. The combination as set forth inclaim 6 wherein said means for establishing a hardlock mode comprisesestablishing a new desired slip relationship at the instant hardlockmode begins by modifying actual slip at that instant to a calculatedamount less than actual slip, said actual slip thereby decaying untilactual slip is zero.
 8. The combination as set forth in claim 4 whereinthe assembly comprises a driving member, a driven member, each memberhaving a friction surface thereon engageable by the other;fluid pressureoperated clutch actuator means for frictionally connecting said drivingand driven members; a pressure source, passage means defined in part bya clutch control valve for distributing control pressure to saidactuator means; electronically operated valve means for establishing aclutch control valve actuating pressure; an electronic processor meansfor varying said actuating pressure according to variations in operatingconditions including torque delivered by said clutch assembly and speedof one of said members; said processor including means for storing atarget slip for said clutch assembly for each set of values of saidoperating conditions, means for measuring actual slip by determiningspeed ratio of the speeds of said members and means for determining acalculated desired slip of said clutch assembly that is less than theactual slip at successive time intervals during clutch actuation wherebya slip decay occurs until said target slip is reached, said desired slipbeing equal to actual slip multiplied by a modifier, and means forestablishing a hardlock mode after said actual slip reaches said targetslip whereby said clutch assembly is fully engaged.
 9. The combinationas set forth in claim 4 wherein said transmission has gearing withmultiple gear ratios, said turbine being connected to driven portions ofsaid gearing, means for detecting the driven speed of said drivenportion, means for continuously monitoring speed ratio of said converteras a function of gear ratio, said detected driven speed and said enginespeed;the monitored speed ratio decreasing upon a command of a shift toa higher speed gear ratio, means for detecting a delta speed ratiofollowing said shift command to provide an indication of the time aratio change begins following said shift command; means for modifyingsaid clutching capacity to effect a decrease in clutch capacity and anincrease in actual slip driving said ratio change; and means responsiveto the attainment of an increase in said delta speed ratio that issufficient to indicate the end of said ratio change whereby saidclutching capacity controlling means resumes its clutch control.
 10. Afluid pressure operated clutch assembly comprising:a driving member, adriven member, each member having a friction surface thereon engageableby the other; fluid pressure operated clutch actuator means forfrictionally connecting said driving and driven members; a pressuresource, passage means defined in part by a clutch control valve fordistributing control pressure to said actuator means; electronicallyoperated valve means for establishing a clutch control valve actuatingpressure; an electronic processor means for varying said actuatingpressure according to variations in operating conditions includingtorque delivered by said clutch assembly and speed of one of saidmembers; said processor including means for storing a target slip forsaid clutch assembly for each set of values of said operatingconditions, means for measuring actual slip by determining speed ratioof the speeds of said members and means for determining a calculateddesired slip of said clutch assembly that is less than the actual slipat successive time intervals during clutch actuation whereby a slipdecay occurs until said target slip is reached, said desired slip beingequal to actual slip multiplied by a modifier, and means forestablishing a hardlock mode after said actual slip reaches said targetslip whereby said clutch assembly is fully engaged.
 11. The transmissionas set forth in claim 4 wherein said transmission has gearing withmultiple gear ratios, said turbine being connected to driven portions ofsaid gearing, means for detecting the driven speed of said drivenportion, means for continuously monitoring the speed ratio of saidconverter as a function of gear ratio, said detected driven speed andsaid engine speed;the monitored speed ratio decreasing upon a command ofa shift to a higher speed gear ratio, means for detecting a delta speedratio following said shift command to provide an indication of the timea ratio change begins following said shift command; means for modifyingsaid clutching capacity to effect a decrease in clutch capacity and anincrease in actual slip during said ratio change; and means responsiveto the attainment of an increase in said delta speed ratio that issufficient to indicate the end of said ratio change whereby saidclutching capacity controlling means resumes its clutch control.
 12. Atorque converter bypass clutch assembly comprising a clutch plate, animpeller housing enclosing a turbine and an impeller, said clutch platedefining with said impeller housing a control pressure chamber, saidplate being actuated by fluid pressure within said impeller housing withthe pressure in said pressure chamber opposing the fluid pressureactuating said plate;valve means for controlling said control pressurein accordance with variations in impeller torque and turbine speedwhereby a controlled slip of said clutch is achieved; means forestablishing a stroke mode pressure to actuate said clutch to apre-engaged condition whereby further pressure increase will causeclutch engagement; means for determining a target slip of said clutchfor each set of turbine speed and impeller torque values; means fordetermining a desired slip; means for effecting a decay in the actualslip of said clutch toward said target slip during operation of saidclutch assembly in an engage mode, said decay being a function of sliperror between said actual slip and said desired slip; and means forestablishing a hardlock mode after said target slip is reached wherebysaid clutch assembly is fully engaged.
 13. The combination as set forthin claim 12 wherein the assembly comprises a clutch plate, an impellerhousing enclosing a turbine and an impeller, said clutch plate definingwith said impeller housing a control pressure chamber, said plate beingactuated by fluid pressure within said impeller housing with thepressure in said pressure chamber opposing the fluid pressure actuatingsaid plate;valve means for controlling said control pressure inaccordance with variations in impeller torque and turbine speed wherebya controlled slip of said clutch is achieved; means for establishing astroke mode pressure to actuate said clutch to a pre-engaged conditionwhereby further pressure increase will cause clutch engagement; meansfor determining a target slip of said clutch for each set of turbinespeed and impeller torque values; means for determining a desired slip;means for effecting a decay in the actual slip of said clutch towardsaid target slip during operation of said clutch assembly in an engagemode; and means for establishing a hardlock mode after said target slipis reached whereby said clutch assembly is fully engaged.
 14. Thecontrol system as set forth in claim 12 wherein said transmission hasgearing with multiple gear ratios, said turbine being connected todriven portions of said gearing, means for detecting the driven speed ofsaid driven portion, means for continuously monitoring speed ratio ofsaid converter as a function of gear ratio, said detected driven speedand said engine speed;the monitored speed ratio decreasing upon acommand of a shift to a higher speed gear ratio, means for detecting adelta speed ratio following said shift command to provide an indicationof the time a ratio change begins following said shift command; meansfor modifying said clutching capacity to effect a decrease in clutchcapacity and an increase in actual slip during said ratio change; andmeans responsive to the attainment of an increase in said delta speedratio that is sufficient to indicate the end of said ratio changewhereby said clutching capacity controlling means resumes its clutchcontrol.